Self-Lubricating Bearings
Development of Test Method for Evaluation of Ageing Mechanisms
Josef Elhag 2015
Master of Science in Engineering Technology Mechanical Engineering
Preface
This thesis is the result of a work that has been carried out at the Divison of Machine Elements at Luleå University of Technology, in collaboration with the Swedish Hydropower Center (SVC) during the period January-May 2015.
I want to express my grateful acknowledgment to my supervisor Dr. Kim Berglund for all the fruitful discussions and the guidance throughout the work.
Your support has been invaluable. To my colleagues at the Division of Ma- chine Elements: thank you for making my time there enjoyable. Lastly, I want to thank my family and friends for the support I always receive from you.
/Josef Elhag Luleå, June, 2015
Abstract
Increased environmental awareness has led to a transition from oil-filled bear- ings to oil-free bearings in hydropower applications, where the oil-free bear- ings have self-lubricating properties. However, there is today a lack of knowl- edge regarding the ageing of these bearing types. Ageing in this context means how the service life is affected by operating parameters such as applied pressure, sliding speed and temperature that in turn controls the friction and wear characteristics of the tribological system. In this thesis, a test method was developed in order to evaluate which ageing mechanisms that are important for self-lubricating bearings.
Previous tests performed on self-lubricating bearings have been conducted by component testing in form of journal bearing rigs. These may however impose unwanted effects in terms of clearance – leading to high peak forces and alternating contact zones between different test samples. A pre-study was therefore conducted with a linear reciprocating test rig, where the results showed that a model test with block-on-block configuration is able to repre- sent tests performed in a journal bearing rig. The block-on-block configuration is beneficial in terms of consistency as it bypasses the issues related to clear- ance and is thus able to perform tests with a conformal contact. However, limitations in terms of e.g. loading capacity and modification difficulties of the rig selected for the pre-study led to the decision to develop a new test rig.
The developed test rig is a linear reciprocating machine, loaded by a hy- draulic actuator and driven by a ball screw unit attached to a rail/wagon en- tity. The performance and properties of the developed test rig allows for inves- tigation of which ageing mechanisms that are important for self-lubricating bearings. Possible measurable parameters are: friction coefficient, wear depth, sliding speed, amplitude of movement and acoustic frequency. It is also pos- sible to measure and control the surface temperature of the counter surface.
The test rig enables testing under relatively mild load cases as well as harsh edge loading situations at variable movement sequences – during operating conditions encountered in hydropower turbines.
CONTENTS CONTENTS
Contents
1 Introduction 1
1.1 Background . . . . 1
1.2 Objectives . . . . 3
2 Theory 5 2.1 Tribology . . . . 5
2.1.1 Lubrication . . . . 6
2.1.2 Wear . . . . 7
2.2 Self-lubricating bearings . . . . 8
2.2.1 Transfer film . . . . 9
2.2.2 Application in hydropower turbines . . . . 10
2.2.3 PV-value . . . . 11
2.2.4 Geometrical design parameters . . . . 12
3 Pre-Study 15 3.1 Selection of test rig . . . . 15
3.2 Friction estimation . . . . 16
3.3 Wear estimation . . . . 17
3.4 Experimental procedure . . . . 18
3.5 Results . . . . 21
3.5.1 Friction and wear results - effect of sliding speed . . . . . 21
3.5.2 Wear of the bearing test specimen . . . . 21
3.6 Discussion & Conclusions . . . . 23
4 Development of Test Rig 25 4.1 Benchmarking . . . . 25
4.2 Requirement specification . . . . 28
4.3 Concept generation - linear movement . . . . 29
4.4 Concept screening - linear movement . . . . 32
4.5 Concept generation - force actuation . . . . 32
4.6 Concept screening - force actuation . . . . 33
4.7 Concept generation - measurement & control . . . . 33
4.8 Concept screening - measurement & control . . . . 37
4.9 Concept selection . . . . 37
4.9.1 Linear movement . . . . 37
4.9.2 Force actuation . . . . 38
4.9.3 Measurement & control . . . . 38
4.10 Detail design - linear movement . . . . 38
4.10.1 Required torque for the motor unit . . . . 41
4.10.2 Structural analysis . . . . 42
4.11 Detail design - force actuation . . . . 44
4.11.1 Structural analysis . . . . 46
4.12 Detail design - measurement & control . . . . 49
5 Discussion 53
6 Conclusions 55
7 Future Work 57
References 61
1 INTRODUCTION
1 Introduction
Globally, renewable energy is growing when environmental demands are in- creasing as a result of raised energy policy objectives due to pollution gen- erated by conventional energy resources. Renewable energy, in form of hy- dropower, is a great power asset as it contributes to approximately 45 % of Sweden’s total electricity production [1, 2]. Hydropower is very beneficial due to its ability and flexibility to operate as the main source of power or as regulating power. It is also advantageous from an environmental perspective, where the operation of hydropower plants do not cause any emissions of car- bon dioxide whilst the water returns to the river after passing through the plant.
Hydroelectricity, which is electricity generated by hydropower, is acquired by conversion of potential energy to kinematic energy which in turn is con- verted to electrical energy. This is obtained through a process where water stored at an elevated level falls down and drives a turbine that is connected to a generator.
Two types of turbines are used in hydropower plants today: reaction and impulse turbines. Reaction turbines spins as the fluid pushes through and past the blades and are categorized as Francis and Kaplan turbines. Impulse turbines spins as fluid with high velocity is fired through nozzles at the tur- bine blades and are characterized as Pelton turbines. The main difference be- tween these types is that the rotors of the Francis and Pelton turbines have fixed blades whereas Kaplan turbines utilize adjustable blades, which allows for high efficiency operation during altered flow conditions [3]. The choice of a turbine is mainly governed by the accessible fall height of the dam, i.e. the head. In Sweden, Kaplan turbines are very common due to low heads (usually around 4-75 m).
1.1 Background
Historically, turbine elements such as runner hubs (which comprises the blades of a turbine) and wicket gates (regulates incoming flow) have been oil-filled to enable lubrication of the moving parts. Environmental legislation have led to a transition from oil-filled components to oil-free components. In practi- cal terms, it means that commonly used materials as e.g. bronze bearings lubricated with oil is to some extent being replaced by bearings with self- lubricating properties to eliminate the need for oil and to reduce pollution [4].
Hydropower offers the stability that wind, wave and solar power does not. As a consequence of this, higher demands are set on hydropower to ef- fectively regulate the power grid in order to meet the market’s requirement of electricity. The primary and secondary regulation have led to increased start
& stops, with elevated transient stresses and altered operating conditions as a
result [5]. These factors, together with the transition to more environmentally friendly bearing materials have opened up a new field of research. There is currently a lack of knowledge about the ageing of these bearings types and how the ageing mechanisms can be evaluated.
Studies have been carried out by Ukonsaari & Prakash [6], Pereira et al.
[7] among others on self-lubricating bearing materials used within the hy- dropower industry. The tests have been conducted in journal bearing test rigs during actual operating conditions encountered in guide vane (wicket gate) and turbine blade bearings; namely at high pressures and low sliding veloci- ties during certain angular movement sequences.
A commonly encountered problem for Kaplan runner hub bearings is edge loading [5], illustrated in Figure 1. This is mainly attributed to the loading sit- uation that causes shaft misalignment, i.e. the bearings are not concentric with the shaft line. Guide vanes may also experience edge loading problems due to shaft misalignment, mainly governed by inadequate installation or insuf- ficient shaft stiffness [8]. A case study performed by Ren and Muschta [8]
illustrates that a misalignment of 0.0023 mm/mm can cause resulting edge pressures to be as large as four to fifteen times the design pressure, depending on the selected bearing material. Tolerance problems leading to undesirable clearance between the shaft-bearing interface may also cause problems; and as a result, only a small area of the bearing may be in contact with the shaft.
These phenomenas cause elevated normal forces and peak pressures, which subsequently accelerates the wear and fatigue in the zones that are heavily loaded [5][9]. When bearings are to be tested, it cannot be assured that the zones would be the same due to the differences between tested bearings in terms of clearance and geometrical properties.
Figure 1: Schematic of Kaplan runner hub - edge loading situation.
1 INTRODUCTION 1.2 Objectives
1.2 Objectives
The main objective with this thesis is to develop a test method to evaluate which ageing mechanisms that are important for self-lubricating bearings in hydropower applications.
The aim of the thesis is to:
• Develop a method to evaluate the ageing of self-lubricating materials used in applications such as trunnion, linkage and wicket gate bearings.
Tribological parameters shall be evaluated in terms of friction and wear characteristics.
• This means detail design of a new test rig, or alternative modification of existing equipment. The design proposal shall include a structural layout, sensors and selection of machine elements such as motors, gear- boxes, bearings etc. Additional sensors for evaluation of ageing mecha- nisms shall also have the ability to be connected if necessary.
2 THEORY
2 Theory
In order to develop a test method to evaluate ageing mechanisms of self- lubricating bearings in hydropower applications, it is essential to have an understanding about tribology in general. This includes knowledge of lu- brication, friction and wear as well as knowledge of mechanical properties of self-lubricating bearings. It is also necessary to have an understanding of how these mechanisms can be measured.
2.1 Tribology
The word tribology originates from the Greek word tribos, which means "I rub" or "to rub". It is often spoken of as: "the science and technology of inter- acting surfaces in relative motion, including lubrication and wear" [10]. Tribol- ogy is a multi-disciplinary science covering subjects from mathematics, fluid mechanics, chemistry, physics and material science to solid mechanics.
The friction and wear heavily depends on the tribological system at hand.
A tribological system involves four elements which contributes to the behav- ior of the system. The first two elements are the materials of the mating sur- faces. The third element is the interfacial medium which is referred to as the lubricant; this element can be in either solid, liquid or gas form depending on the application. The fourth element is the surrounding element which is commonly air or some other fluid. The bulk of the information presented in this tribology section is gathered from Van Beek [11]. An illustration of a tri- bological system, containing the four tribological elements, is shown in Figure 2 below.
Figure 2: Tribological system.
2.1.1 Lubrication
The lubricant is a fundamental and necessary element in a tribological sys- tem. By adding a lubricant between two surfaces, the friction and wear may be reduced significantly. A lubricant can, depending on the application, carry the subjected load by formation of a pressurized (hydrodynamic) film gener- ated by relative motion and a converging gap. Lubrication is often divided into three main regimes: boundary, mixed and full film lubrication. Differ- ent applications operates in different regimes; e.g. thrust bearings located in hydropower plants operates in the full film regime where the surfaces are completely separated in order to minimize the friction and wear of the thrust halves. Piston-cylinder liner is a typical example of an application that works in between these three regimes, due to that the velocity is zero in the top and bottom dead centers of the piston stroke. The lubrication regime mainly de- pends on the relative motion between the surfaces, together with the viscosity of the lubricant and the applied load. The Stribeck curve, shown in Figure 3, is widely used in Tribology to define these regimes. A brief explanation of the different lubrication regimes, located on the Stribeck curve, is transcribed below.
Figure 3: Stribeck curve.
Boundary lubrication
If the load is mainly transmitted by mechanical contact, the lubrication regime is defined as boundary lubrication. This is acquired at low velocities and at high loads. Substantial asperity contact occurs where high friction and wear is a result of this action. The lubricant’s main function in this regime is to control friction and wear between the mating surfaces.
2 THEORY 2.1 Tribology
Mixed lubrication
In this lubrication regime, the load is partly transmitted by mechanical con- tact and partly by hydrodynamic pressure as the sliding speed is increased (or load decreased). The friction and wear in this regime decreases as more hydrodynamic pressure is built up.
Full film lubrication
A further increase of the sliding speed may build up enough hydrodynamic pressure to separate the surfaces, hence full film lubrication is acquired. The friction and wear in this regime is greatly reduced since it is the lubricant in between the surfaces that is sheared.
2.1.2 Wear
Wear is defined as removal of material from surfaces, quantified by mass or volumetric loss. It has a great effect on the lifetime of machine elements. Cer- tain applications are designed to wear, such as wear plates found in iron ore crushers or manufacturing processes as e.g. milling where controlled wear takes place. Other applications are highly vulnerable to wear, for instance shaft–bushing interfaces in hydropower turbines. Uncontrolled wear in such applications can cause undesired clearance which in turn might lead to seal failure or vibrations in the system. Wear is commonly classed by four differ- ent mechanisms: abrasive, adhesive, corrosive (chemical) and surface fatigue wear.
Adhesive wear
Adhesive wear occurs when strong attractive adhesive bonds between asper- ities cause micro-welding. High levels of heat and friction develops during the interaction. This action can result in transfer of material from one surface to another, either temporarily or permanently, due to that the bond is stronger than the base material. In the case of temporary adhesion, free wear particles are often the result which in turn may cause severe damage to the machine.
A greater risk of adhesive wear is imposed if the two surfaces in contact have compatible materials, e.g. steel on steel. It is common that dissimilar materials are used to prevent this type of wear mechanism.
Abrasive wear
The harder surface wears the softer surface by scratching of material. This process is enhanced if the difference in hardness is large between the mating surfaces and if the harder surface is rougher. A common, but controlled, pro- cess where this action takes places is e.g. in the wood industry when abrasive paper wears down wood in a planing machine. Abrasive wear is commonly split-up into two mechanisms: two-body and three-body abrasion.
• Two-body abrasion: Applications encountering this type of wear gen- erally experiences more material loss than in the case for three-body abrasive wear. Characteristic for this wear mechanism is that the wear markings are parallel to the sliding direction.
• Three-body abrasion: Hard abrasive particles in between the surfaces (originating from either one of the surfaces in contact or from external contamination) plastically deforms the softer surface.
Surface fatigue wear
Surface fatigue wear is initiated by fatigue-related (sub-surface) cracks, trig- gered by repeated stress cycles. This form of wear commonly takes place in tribological systems involving rolling contacts combined with some slip, e.g.
rail and rolling element bearing applications. When large particles of material breaks away from the surface, the phenomena is called pitting. If pits are cre- ated in a smaller order of magnitude, the phenomena is denoted spalling.
Corrosive wear
Arises as a result of formation of reaction products in a corrosive environment (corrosive liquids or gases). These reaction products strongly affects the wear process. This type of mechanism is commonly referred to as tribochemical wear and can contribute to excessive wear during e.g. oxidation at raised temperatures.
2.2 Self-lubricating bearings
The mechanical and tribological performance of plastic bearings can be sig- nificantly improved if reinforcement is carried out with fibers and with the addition of solid lubricant fillers. Fibers have the ability to increase the me- chanical strength of the bulk and thereby affect the limiting PV-value together with the friction and wear characteristics. Solid lubricants distributed in the material are able to transfer onto the mating surface during sliding motion, thus forming a lubricating (transfer) film that results in less friction and wear.
Common encountered fibers and dry lubricant fillers are presented in the list below [11][12]:
• Polyfluorotetraethylene (PTFE): The transfer film that adheres to the mat- ing surface has low shear strength, i.e. the macro molecules slip easily along each other, which contributes to reduced friction and wear. Oper- ating temperatures are restricted to around 260◦C.
• Graphite: Water vapor is required for graphite lubrication, hence it is suitable and effective where moist and aqueous environments are present.
Operating temperatures are up to 450◦C. It is commonly used in applica- tions where minor lubricity is needed and a thermal insulating coating is required.
2 THEORY 2.2 Self-lubricating bearings
• Molybdenum disulfide (MoS2): Generally easy to shear. Has a wide range of application areas and works both in dry and moist environ- ments, contrary to graphite.
• Glass fiber (GF): Improves the mechanical strength and creep resistance.
However, the fiber ends in particular may increase wear of the counter surface which leads to increased friction.
• Carbon fiber (CF): Improves the mechanical strength, heat conductivity and PV-value. Fibers of carbon are less abrasive than fibers of glass.
• Aramid fiber (AF): Does not exhibit as good reinforcing effects as glass or carbon fiber, due to lower fiber strength. This fiber however shows good properties in terms of resistance to abrasion, particularly for counter sur- faces with lower hardness as e.g. bronze, aluminum or plastics.
2.2.1 Transfer film
During relative motion between two surfaces, e.g. polymer on steel, particles from the polymer are torn from the surface due to local shear stresses and by softening of regions of the polymer surface caused by heating. If the torn particles adheres to the counter surface’s asperities, a transfer film will begin to form. This transfer film will continue to develop as more particles adhere to the counter surface [13, 14].
Studies conducted by Hollander and Lancaster [15] in the early 70’s showed that the formation of a transfer film, with PTFE and other polymers sliding against metal counter surfaces, reduced the surface roughness of the latter.
The authors proposed that the film formation led to an increase of the true area of contact which thereby reduced the local stresses in the contact and thus decreased the wear rate of the system. The authors also reported that the situation of hard rigid counter surface asperities, penetrating the softer polymer, no longer exists when a continuous transfer film is formed.
Rhee and Ludema [13] showed that the formation of a transfer film, with Delrin (POM-based), PTFE filled Delrin and Nylon, against steel and quartz counter surfaces was highly affected by the counter surface roughness and the applied load. They reported that the transfer film formation generally was promoted if the applied load and roughness were increased.
Formation of a transfer film is thus dependent upon several parameters that affect the tribological performance. The sliding speed together with the operating temperature and the mating material also have a great importance regarding the formation as they contribute heavily to the friction and wear of the tribological system.
The authors also proposed that initially, at low sliding speeds, the trans- fer film is not continuous and allows for “rolling” of semi-solid particles. The friction in this stage can however be considered to be low due to the rolling
action taking place. When the sliding speed increases, the coefficient of fric- tion increases due to heating of the transfer film which causes it to behave more as a viscous fluid, thereby increasing the contact area. However, the increase of the friction coefficient is nearly independent of the sliding speed when enough heat has been generated to obtain a continuous transfer film.
When a continuous film is acquired, the wear is generally minimal.
Brainard and Buckley [16] even observed transfer of PTFE to metal counter surfaces in static contacts with the usage of field ion microscopy and Auger emission spectroscopy. They found that temperature and time (i.e. time- dependent creep in the polymer) have great effect upon the force of adhesion which gives rise to transfer film formation. The transfer film formation and its stability also depend upon the cohesive bond strength of the transfer film and the adhesive bond strength between the film and the sliding surfaces [17].
An additional factor that has a significant effect on the tribological behav- ior in terms of transfer film and wear is the influence of fillers. Studies con- ducted by Bahadur [18] have shown that the wear resistance may be enhanced or degraded if certain fillers are added into the material composition. Accord- ing to Bahadur, the increase of wear resistance is mainly due to reaction prod- ucts that are generated which enhances the bonding between the counter sur- face and the transfer film. The decrease of wear resistance is mostly because of discontinuities generated in the material due to certain fillers.
All these parameters are important in the aspect of controlling the wear encountered in a tribological system as they have an important task of pro- tecting the rubbing polymer surface from e.g. harder metal asperities on the counter surface. Wear however usually occurs due to loss of material before transfer has been made onto the counter surface [14].
2.2.2 Application in hydropower turbines
There are numerous reasons why self-lubricating bearings are used in e.g. Ka- plan turbines, especially in the runner hub which has the task of regulating the runner blades. Some of the benefits that can be achieved are listed below [7]:
• Reduced friction between the runner blades and trunnion bearings re- duces the force required to move the blades, thereby posing less de- mands on the actuating system.
• Lower wear rate prolongs the lifetime of the bearings.
• Works sufficient in the boundary lubrication regime (caused by low- oscillating motion and high loads). It is difficult to achieve a sufficient lubricating film during these conditions if oil or grease were used.
• The solid embedded lubricants provide adequate lubrication, even when movement starts during high load operating conditions.
2 THEORY 2.2 Self-lubricating bearings
• Self-lubricating bearings with relatively high Young’s modulus are able to prevent misalignment of the blade trunnion.
• Environmental benefits such as less oil usage and risk of oil leakage by replacing oil-filled runners with oil-free runners.
2.2.3 PV-value
To accommodate for reliable bearing operation, it is important to have knowl- edge about the frictional heat generation in the bearing interface as it is the limiting factor of polymer composite bearings. The temperature build-up is mainly attributable to the operating pressure and velocity (PV) together with other parameters such as the system thermal dissipation and the friction coef- ficient.
Taking all of these parameters into consideration when designing a bearing would be a complex procedure. However, experience has through the years shown that usage of PV-values when designing bearings can be an effective estimation regarding the service life, if used correctly. The basic idea behind this estimation is that the combined pressure and velocity value should not be exceeded during operation. If the limiting PV-value is exceeded, the wear rate is generally accelerated as thermal softening or plastic deformation occurs.
It is therefore of great importance that factors such as peak and shock loads are taken into consideration so that the maximum compressive stress is not exceeded [19]. An illustration of a typical PV-diagram is shown in Figure 4.
Figure 4: Pressure-velocity diagram.
2.2.4 Geometrical design parameters
There are also other important parameters that can have an effect on the tribo- logical behavior, apart from those previously discussed (i.e. mating materials, applied load, temperature, sliding speed etc.). Parameters such as the length- to-diameter ratio (L/D-ratio) and the shape-ratio (S-R) takes the geometrical design into account. It is important to have knowledge about the variations in geometrical design that different manufacturers supply as it enables one to experimentally evaluate the influence of these parameters and thereby deter- mine which configurations that are best suited for certain applications.
L/D-ratio
Below is a summarized list where some of the most common encountered L/D-ratios for self-lubricating bearings in the hydropower industry are pre- sented:
• The recommended L/D-ratios for Thordon Thorplas bearings are maxi- mum 2:1 and minimum 1:1 [20].
• Federal-Mogul Corporation recommends L/D-ratios from 0.5:1 to 1.6:1 [21] for their self-lubricating deva.tex bearings.
• Kamatics Corporation recommends L/D-ratios of less than 1.5:1 for their self-lubricating Karon bearings to keep both edge loading and pin (shaft) bending to a minimum [22].
• Tufcot recommends to keep the L/D-ratio below 2:1 for their bearings, if possible [19].
• Jones et al. [4] recommends L/D-ratios from 1:1 to 2:1 for thick-walled bushings and L/D-ratios ranging from 0.35:1 to 1:0 for thin-walled bush- ings.
In conclusion to the points listed above; the L/D-ratio ranges from 0:35:1 up to 2:1, depending on the manufacturer. This data can be associated with the information described regarding the PV-value. If for instance the calculated pressure value exceeds the acceptable value in a PV-diagram, one could in- crease the diameter or length in order to decrease the pressure (due to an in- crease of the contact area). The same applies to the case when the operating velocity is considered too high; one could then decrease the shaft diameter in order to decrease the tangential velocity [19].
Shape ratio
Compression experiments have shown that the stress-strain curve of a poly- mer is heavily influenced by the geometrical shape of the test specimen. The shape ratio (or shape factor) is commonly used to denote this effect. The shape ratio is defined as the quotient between the loaded area and the area which is
2 THEORY 2.2 Self-lubricating bearings
free to bulge. During loading action, the specimen will deform in accordance to the applied force and in inverse proportion to the shape ratio. There might not be a substantial change in volume, but the shape however might change significantly. The displacement of a polymer decreases, for a given load, if the shape ratio is increased due to a decrease of the bulge area. This in turn yields a higher load carrying capacity. The amount of creep, i.e. time dependent de- formation, can also be reduced if the shape ratio is increased [20]. The shape ratio is thus:
Shape ratio = Loaded area
Area free to bulge (1)
For a journal bearing, the shape ratio is acquired by:
The loaded area can be approximated as the projected area (seen as the blue plane in Figure 5):
Loaded area = d · l (2)
Where d is the diameter and l is the length of the bearing. The area free to bulge is:
Bulge area = 2 · d · t (3)
The constant 2 is due to that both sides of the sleeve may bulge during loading and the parameter t is the bearing thickness. Equation 2 and 3 yields:
Shape ratio = d · l 2 · d · t = l
2 · t (4)
Figure 5: Cross-cut of a journal bearing.
For a rectangular/quadratic test specimen, the shape ratio is acquired by:
The loaded area (seen as the gray plane in Figure 6) is:
Loaded area = x · y (5)
Where x and y are the length and width of the specimen. The area which is free to bulge is comprised of the sides of the specimen, i.e.:
B1= z · y (6)
And
B2= z · x (7)
If the specimen is quadratic, then B1 = B2. By using Equation 5 to Equation 7, the shape ratio is obtained by:
Shape ratio = x · y 2 · B1+ 2 · B2
= x · y
2 · z · y + 2 · z · x = xy
2z(y + x) (8)
Figure 6: Cuboid test-specimen.
3 PRE-STUDY
3 Pre-Study
A pre-study was conducted with the aim of investigating if it is possible to represent a component test by means of a model test, in order to evaluate which ageing mechanisms that are important for self-lubricating bearings in hydropower applications. Model tests are generally quick, cost effective and advantageous in terms of isolating test parameters under well-controlled test conditions [23].
Component tests of this bearing category have earlier been conducted by Ukonsaari & Prakash [6] among others. The tests have been performed in journal bearing test rigs during high pressure and low velocity oscillating load cycles. However, as mentioned in the background section, component tests in journal bearing rigs may cause effects such as undesirable clearance between the shaft-bearing interface – leading to non-uniform contact zones and high peak forces. The contact zones may also vary between different tests as each specimen is not likely to be identical in terms of geometrical properties. If a new material is tested, it may be difficult to distinguish if the results depend on the geometric characteristics or the properties of the material. This feature is illustrated in Figure 7, where the contact zones are shown as the red lines.
The selected model test method which was thought to represent a journal bearing application in a thorough manner was a block-on-block configuration.
This configuration was considered to be capable of simulating a journal bear- ing by means of a conformal contact, thereby removing issues with clearance related to high peak forces, bearing replacement and various manufacturing processes. The results from the performed tests were then compared with results obtained by Ukonsaari & Prakash [6] in terms of friction and wear characteristics, in order to determine if the selected test configuration is repre- sentative.
Figure 7: Schematic of various clearance cases.
3.1 Selection of test rig
The test rig that was used to conduct the experiments was a TE-77 Cameron- Plint, with a block-on-block configuration. The rig works in a manner where a vertical force is applied onto a test specimen by a loading spring with the ca- pacity of 500 N. A linear, reciprocating movement of the loaded test specimen
is then achieved by an eccentric mechanism driven by an electric motor. An illustration of the test setup is shown in Figure 8.
(a) TE-77 Cameron-Plint test rig. (b) Block-on-block configuration.
Figure 8: Test rig configuration.
3.2 Friction estimation
Friction coefficients for the conducted tests were obtained by the horizontal force, measured by a load cell, divided by the vertical force applied by the loading spring. Mathematically, the friction force is expressed as:
Ff = µN (9)
Where Ffis the friction (horizontal) force, µ the friction coefficient and N the normal (vertical) force. Dividing both sides of Equation 9 with N yields the friction coefficient:
µ =Ff
N (10)
A typical stream of friction coefficient data, obtained with the test rig, is shown in Figure 9.
3 PRE-STUDY 3.3 Wear estimation
Figure 9: Friction data stream.
The mean friction coefficient is obtained by summing all the friction data val- ues and then dividing the summed value with the number of data points.
3.3 Wear estimation
Wear measurement in the rig was carried out with the usage of an LVDT sen- sor that measures displacement, mounted on top of the specimen holder. The dimensions of the specimens together with the weight were also measured before and after each experiment to have another independent source of wear encountered during the tests. Pre-weight and height measurements are how- ever not as reliable, in terms of steady-state wear, as measurements with the LVDT sensor since it is difficult to estimate the amount of wear that transpires during the running-in sequence of each test.
The amount of wear in a tribological context is commonly acquired through the use of Archard’s wear equation which can be written as [24]:
h = kps (11)
Where h is the wear depth, k the wear coefficient, p the applied pressure and sthe total sliding distance. The applied pressure is acquired by the applied force, F , divided by the contact area A, i.e.:
p = F
A (12)
The total sliding distance is obtained by the average sliding velocity, v, times the elapsed time t:
s = vt (13)
For a linear reciprocating motion, the average sliding velocity is obtained by:
v = 2f l (14)
In this expression, f is the reciprocating frequency and l is the stroke length.
Substituting Equation 14 into Equation 13 and finally Equation 12 into Equa- tion 11 yields the wear depth:
h = 2kf ltF
A (15)
Solving for the wear coefficient, k, yields:
k = hA 2F f lt = h
ps (16)
In Figure 10, the wear depth is plotted against the product of applied pressure and sliding distance. This figure shows typical raw data acquired from the conducted experiments by the LVDT sensor. The mean wear coefficient is attained by linear least-square fitting of the data points.
Figure 10: Wear depth vs. pressure times distance.
3.4 Experimental procedure
Before each test was conducted, the specimen and counter surface were cleaned with ethyl alcohol and dried with a tissue to ensure that contaminants were not present. Each experiment started with a run-in sequence of 15-25 min- utes, depending on the sliding speed. Reciprocating frequencies below 2 Hz, i.e. the minimum frequency acquired by the rig, were obtained by usage of a gearbox with a ratio of 20:1. The duration of the gearbox tests were scaled so that sliding distances of similar magnitude were obtained as for the case with frequencies of 2 Hz. Every test was performed under dry lubricated con- ditions and each specimen had equal geometrical dimensions of 4x4x4 mm.
3 PRE-STUDY 3.4 Experimental procedure
This means that the length-to-diameter ratio was equal to 1 and the shape ratio roughly 0.4, according to Equation 8 (the shape-ratio is higher than the theoretical value due to that approximately 1.5 mm in the vertical direction is required to clamp the specimen in the holder). The tests were conducted at contact pressures of 25 MPa, as higher were not possible due to limited load- ing capacity of the test rig.
Specimen and counter surface topography were also measured before and after the tests with a Wyco NT1100 optical profilometer. Counter surfaces used in the tests were made of steel, with a surface finish Ra roughly equal to 0.3 µm. The experiments were performed during the operating conditions presented in Table 1.
Table 1: Experimental operating conditions.
Contact pressure [MPa]
Temperature [◦C ]
Sliding velocity [mm/s]
Stroke length [mm]
Time [h]
25 Room 36.8 8 13
25 Room 14.4 8 39
3 PRE-STUDY 3.5 Results
3.5 Results
3.5.1 Friction and wear results - effect of sliding speed
The acquired mean friction and wear coefficients from the experimental mea- surements are presented in Figure 11 below.
(a) Friction results. (b) Wear results.
Figure 11: Friction & wear results for Thordon Thorplas material.
The obtained mean friction coefficient for the tests at 36.8 mm/s was 0.095 with a slight deviation of around 0.002. For the tests conducted with the gear- box at 14.4 mm/s, the mean friction coefficient was a lower value of 0.069 with a minor deviation of 0.001.
The obtained wear data pointed towards a trend where lower wear oc- curred at higher sliding velocities. The mean wear coefficient was 6.50 · 10−10 mm2/Nwith a lower limit of 1.45 · 10−10and an upper limit of 2.45 · 10−10for the tests conducted without the gearbox. For the gearbox tests, the mean wear coefficient was at 12.0 · 10−10mm2/Nwith a deviation of 1.32 · 10−10.
3.5.2 Wear of the bearing test specimen
An illustration of a test specimen surface measured with the Wyco NT1100 optical profilometer, before and after testing, is shown in Figure 12 and Figure 13.
Figure 12: Unworn Thordon Thorplas bearing test specimen.
From this figure, one can identify a wavy pattern indicated by the peaks and valleys which have a noticeable height difference. The wavy pattern is ob- tained from a milling manufacturing process that has been carried out be- forehand. Another apparent feature is that the test specimen has a slightly curved profile, which is reflected by the height difference between the edges (green/red color) and the middle region (blue color).
Figure 13: Worn Thordon Thorplas bearing test specimen.
This figure shows the specimen after testing has been conducted; it demon- strates a fairly worn and smoothed surface, which indicates that the initial peaks have been evenly abraded throughout the test. The curved profile is also apparent in this figure. There might however be some noise in the mea- surement which can be attributed to the small asperity peaks in the middle region of the surface.
3 PRE-STUDY 3.6 Discussion & Conclusions
3.6 Discussion & Conclusions
Several conclusions can be drawn from the obtained experimental results re- garding the friction and wear characteristics of the bearing material Thordon Thorplas. The results showed that a lower friction coefficient was obtained when the sliding velocity was decreased whilst a higher wear coefficient was acquired. The performed tests were similar in terms of operating conditions and the parameter which was varied was the sliding speed (approximately 2.5 times difference). It is therefore possible that the differentiating factor was that more frictional heat was generated during the tests performed at higher sliding velocity.
Yang et al. [25] studied the effect of temperature for self-lubricating PTFE sliding against stainless steel and observed that more transfer occurred at raised temperatures. In conclusion to the result obtained in this thesis; a pos- sible explanation and an assumption is that more transfer transpired and that the transfer film-counter surface adhesion was stronger in the experiments carried out with higher sliding velocity. As a result of this, stronger adhesion may also have occurred between the transfer film and the bearing test speci- men – which in turn lead to an increase in friction. This friction-velocity ten- dency has earlier been observed and is demonstrated in a review conducted by Myshkin et al. [26]. The transfer film may also have acted as a more protec- tive layer for the counter surface, hence lowering the wear rate.
Another important conclusion is that the acquired mean wear coefficients, with the block-on-block configuration, are within the same order of magni- tude as those obtained by Ukonsaari and Prakash [6] in a journal bearing test rig, i.e. within a factor of 10−10. The obtained wear coefficients were neverthe- less slightly higher at the operating pressure of 25 MPa. It is worth mentioning that the wear coefficients lie close to the coefficients acquired by the authors at 50-65 MPa.
It is possible that these results were influenced by the initial curved sur- face shape of the bearing test specimens and the patterns obtained from the milling manufacturing process seen in Figure 12. The geometrical shape could have caused a reduction of the apparent contact area and as a result lead to an increase in contact pressure. The sliding speed may also have had an impact upon the variance of the results with regard to the tests conducted by Ukon- saari and Prakash [6], as those were conducted at a lower sliding velocity.
The main objective of these measurements was to determine if the se- lected test configuration, i.e. the block-on-block, is suitable for testing of self- lubricating bearings in hydropower turbines. The acquired results confirms this hypothesis and it can therefore be concluded that the selected test config- uration is valid.
However, the TE-77 Cameron-Plint test rig has some limitations such as the load capacity which is essential when evaluating various bearing materi- als. Another constraint is that the amplitude of the motion is constant, and
therefore restricted, due to that the linear reciprocating motion is achieved by an eccentric mechanism. To modify this rig in order to make it more flexi- ble, one would have to redesign the whole application. This is not considered worthwhile since the platform the rig is built on does not allow any major modifications. This analysis led to the decision to develop an entirely new test rig, with block-on-block configuration, in order to assess ageing mecha- nisms of self-lubricating bearings used in hydropower applications.
4 DEVELOPMENT...
4 Development of Test Rig
This chapter of the thesis includes the development of the test rig. The devel- opment process was inspired by Ulrich and Eppinger [27] and followed the steps according to:
• Benchmarking
• Establishment of requirement specification & system decomposition
• Concept generation
• Concept screening
• Concept selection
• Detail design
4.1 Benchmarking
Existing rigs, used for testing of self-lubricating bearings in hydropower tur- bines, were investigated to gain knowledge of operating parameters and test configurations. It was found that journal bearing test rigs are mainly used to evaluate these bearing types. Some of these rigs, together with their operating capabilities, are presented in this section.
Pereira et al. [7] conducted tests on the self-lubricating bearing material deva.bm, which is commonly used for blade bearings in Kaplan runner hubs.
The test procedure included a loop of 60000 cycles with an operating pattern according to:
• 100 cycles of ±45◦movement
• A pause of 10 minutes
• 1000 cycles of ±5◦movement
• A pause of 10 minutes
• 2500 cycles of ±3◦movement
The authors performed experiments with a contact pressure of 40 MPa and a sliding speed of 0.01 m/s. The test rig that was used is presented in Figure 14.
Figure 14: Journal bearing test rig for evaluation of Kaplan blade trunnion bearings [7].
Journal bearing tests with the aim of simulating Kaplan runner and guide vane bearings have also been performed by Demianov et al. [28] with the usage of a similar test rig, property of Power Machines/LMZ. The test rig, shown in Figure 15, has the capacity of operating during the following condi- tions:
• The shaft is able to rotate 45◦in both directions, relative to the bearings.
• The bearings may be loaded with pressures ranging up to 75 MPa.
• It is possible to acquire sliding velocities up to 4 mm/s.
Figure 15: Journal bearing test rig for evaluation of Kaplan blade trunnion and guide vane bearings [28].
In a technical report written by the U.S Army Corps of Engineers [4], the au- thors presents testing equipment for evaluation of self-lubricating bearings in
4 DEVELOPMENT... 4.1 Benchmarking
hydropower applications. The setup is shown in Figure 16. A standardized friction and accelerated wear test is also proposed:
• Minor oscillating movement: ±1◦continuously at 2 Hz
• Major swings: ±15◦once every 15 minutes
• Applied static load: sufficient to provide 3300 psi (23 MPa)
• Superimposed load: should be able to superimpose variable bearing load of ±1000 psi on the 3300 psi test load
Figure 16: Greaseless bushing test apparatus [4].
Other test rigs that have been used to study tribological characteristics of bear- ing materials encountered in hydropower applications are found in e.g. [29]
and [30].
4.2 Requirement specification
A number of important parameters were acquired from the pre-study, bench- marking work and the theory part of the thesis. These parameters, together with additional factors considered necessary to evaluate ageing mechanisms of self-lubricating bearings, have been broken down to a requirement specifi- cation which is presented in Table 2. Each criteria has been assigned as Require- ment or Wish, meaning that that the test rig must fulfill the specified criteria or that it is desirable. Possible criteria have been assigned metrics while the remaining are considered as binary (i.e. yes/no).
Table 2: Test rig - requirement specification.
Criteria Req/Wish Target
Design
Robust Req
Modular Wish
Enable visualization Wish
Ambient isolation Wish
Performance
Enable lubrication Req
Adjustable pressure Req 0-90 MPa
Adjustable amplitude Req 1-100 % of contact length
Adjustable sliding velocity Req 0-20 mm/s
Adjustable specimen L/D-ratio Req 0.35:1 - 2:1
Adjustable specimen shape-ratio Req Minimum of 2 Enable start & stops Req
Short downtime Req Less than 30 minutes
Control surface temperature Wish Measurement capability
Applied force/pressure Req
Sliding velocity Req
Amplitude Req
Temperature Req
Friction coefficient Req
Wear depth Req
Acoustic emission Wish
To fulfill the requirements assigned as targets in Table 2, one must first arbi- trarily specify the geometrical dimensions of the intended test specimens.
Specimen cross-sectional dimensions of 20x10 mm to 30x15 mm were con- sidered to be acceptable as an applied vertical force of 20-40 kN is required to obtain a contact pressure of 90 MPa. Length-to-diameter ratios ranging up to
4 DEVELOPMENT... 4.3 Concept generation - linear movement
2:1 at the desired contact pressure is also possible with this configuration. An important factor associated with this decision is that the intended specimen size shall be larger than the size that the TE-77 Cameron-Plint rig can manage, in order to have the ability of transferring friction and wear characteristics between different geometries.
A decomposition of the test rig’s functions was performed, based on the re- quirement specification, to obtain an easier overview of the intended system.
The test rig, i.e. the main system, was decomposed into two sub-systems:
"Linear movement" and "Force actuation". The first sub-system includes the horizontal reciprocating motion while the second sub-system comprises the vertical force application. These systems are interlinked with an additional system denoted "Measurement and control", due to that measurement of vari- ous quantities connects these systems together. An overview of the decompo- sition is shown in Figure 17.
Figure 17: System decomposition.
4.3 Concept generation - linear movement
In this phase, the aim was to come up with various solutions to how a linear reciprocating movement can be achieved which meets the requirements stated in Table 2.
Actuator driven
This concept is built up by two linear rails and four wagons, where two wag- ons are mounted on top of each rail. The wagons have internally mounted bearings in order to reduce the friction in the rail-wagon interface. A top plate connects the wagons and thus forms a carriage unit. The carriage unit is lin- early driven by an actuator connected to the assembly through an adapter, see Figure 18. This concept can be seen as rather flexible due to that the actuator may be e.g. pneumatic, hydraulic or electric.
Figure 18: Linear movement - actuator concept.
Rack & pinion
The "Rack & pinion"-concept is similar to the "Actuator driven"-concept due to that two linear rails, with two wagons mounted on top of each rail, are con- nected by a top plate. One major difference is that the linear unit is connected to a rack through an adapter. An electric motor powers a pinion via a gear- box and thus enables movement of the rack-linear unit. An illustration of the concept is shown in Figure 19.
Figure 19: Linear movement - rack & pinion concept.
4 DEVELOPMENT... 4.3 Concept generation - linear movement
Ball screw
This concept is also similar to the "Actuator driven"-concept. The difference is that the propulsion of the carriage unit is obtained by a ball screw coupled with a servo motor. The concept is shown in Figure 20.
Figure 20: Linear movement - ball screw concept.
Belt drive
The idea behind this concept is to have a belt drive unit with a carriage at- tached. The belt’s drive side is connected to a gearbox which is driven by an electric motor. An alternative solution is to drive the unit with the usage of a chain. The concept is illustrated in Figure 21.
Figure 21: Linear movement - belt drive concept.
4.4 Concept screening - linear movement
The "Rack & pinion"-concept was excluded because it was thought of having difficulties in achieving a rather small and adjustable amplitude. Since racks have a constant pitch, i.e. the distance between each gear, the amplitude can only be adjusted by quantized increments. If the increments however were small enough, it would probably be rather difficult and expensive to manufac- ture. Reducing the geometrical dimensions also implies that the mechanical strength is affected adversely.
The "Belt drive"-concept was excluded because no suitable unit was found with the capability of withstanding the desired vertical forces and the abil- ity to generate enough horizontal (thrust) forces. Another important factor considered was the accuracy of the motion; rubber belts may experience hys- teresis which thereby affects the positioning accuracy in a negative manner, especially at lower amplitudes. Chains on the other hand are not affected by hysteresis. Instead, backlash might become a position accuracy problem at the dead centers of each stroke. The concepts that made it through the screening stage were thus the "Actuator driven"-concept and the "Ball screw"-concept.
4.5 Concept generation - force actuation
Numerous concepts were generated regarding the force actuation system. The ideas included the following force application solutions:
• Dead weight
• Hydraulic
• Electric
• Pneumatic
• Spring loaded
• Lever arm - with any of the above included
Several challenges were associated with the force application. The fact that a vertical load is applied onto a test specimen while a linear horizontal motion is subjected underneath causes reaction forces in two axes. Most actuators are not designed to withstand radial loads, hence a solution had to be developed to overcome this issue.
The solution to this problem is a design that consists of a housing where an actuator is mounted. The idea is that the actuator exerts a force onto a push cylinder via a pressure plate. The push cylinder is located in a bushing to ac- commodate the radial forces resulting from the linear movement. The push cylinder and housing also has a through hole intended for a locking pin, in
4 DEVELOPMENT... 4.6 Concept screening - force actuation
order to hold the cylinder in place during e.g. sample replacement or mainte- nance work. The concept is shown in Figure 22.
Figure 22: Force actuation concept.
4.6 Concept screening - force actuation
Application of sheer dead weight was omitted since weights in the order of 2-3000 kg are needed to obtain a contact pressure of 90 MPa with the intended specimen size. Loading by means of a pneumatic actuator was also excluded due to difficulties in finding a suitable unit with sufficient load capacity.
The work therefore proceeded with hydraulic and electric actuators whilst a spring loaded application was still considered to be a viable option. A spring loaded solution, with e.g. disc springs (intended for short travel and high forces), has the ability to be combined with various actuators. This solution implies that the working actuator only has to apply an initial force, whereas a break can be applied to hold the disc springs in a compressed state.
4.7 Concept generation - measurement & control
This section is split in two main parts: friction and wear measurement. Other measurable parameters as e.g. temperature are further discussed in the detail design chapter. The same applies to the control unit of the system, i.e. partic- ular measuring devices, control schemes etc.
