Review of Literature

The first designs of these bearings had a plain bore, but this was modified several times in order to improve the operation of the bearing. The most popular design used today is the straight fluted bearing. This design of bearing consists of a number of load-carrying lands or staves, separated by flutes orientated axially in the bearing (See Fig. 13.2). The flutes supply the bearing with lubricant, which enters at one end of the bearing and leaves at the other. The flexible rubber bearing liner is bonded onto the outer rigid shell. Environmental concerns with oil and grease lubricated bearings have brought about an increase in the use of water lubricated bearings.

Journal bearing design and performance has been a major research area for the Tribology Research Group at Manipal Institute of Technology [4-10], and Queensland University of Technology (QUT) for many years [11-17].

Water-based bearing lubrication has been an important theme in this endeavour and has been investigated through undergraduate student projects and national and international industrial consulting. In 1990, a test rig [11] was developed at QUT to assess the wear performance of non-metallic, water lubricated journal bearings. The rig was subsequently enhanced to determine the pressure distribution in the water film and the optimum groove size and number. Conventional practice is to assume a linear pressure distribution along the groove [12, 13]. Subsequent experimental work by Hargreaves et al. [16] has cast this assumption into doubt (Fig. 13.3). Indeed, preliminary Computational Fluid Dynamics (CFD) analysis by Pai et al. [17] suggests that the liquid lubricant flows into the clearance space in the unloaded section of the bearing periphery and out of the clearance in the loaded section. This represents a highly complicated flow phenomenon.

Pai et al. [17] modelled the clearance volume of the water bearing for single and three grooves using CFD to study the fluid flow phenomenon. Their objective of study was to predict the pressure in the lubricant (water), measure the pressure in the lubricant (water) film in the journal bearing operating with various conditions.

Manipal Group Instutuion

The inlet pressure was set to 50 kPa, normal to the boundary and the outlet pressure was set to 42 kPa. The bearing shell was modelled as a moving wall with absolute motion of 0 rad/s. The water in the clearance volume was modelled as type 'fluid' with rotation axis origin and direction same as that of the journal. The segregated solver was used for the solution and the flow was assumed to be laminar and steady. There was a difference in the predicted pressure contours and the experimentally measured contours. Axial pressure variations along the groove agreed with experimental trends from a qualitative perspective.

The experimental measurements of pressure in Fig. 13.3 indicate that the pressure remains constant along the groove and drops sharply at the inlet and exit rather than following a linear distribution along the groove. The groove flow in the unloaded part of the bearing is less significant to the bearing performance because the pressures there are much less than in the loaded part, and therefore have little contribution to load-carrying capacity.

Some research on the groove arrangement in journal bearings has been conducted [18, 19]. They analysed the performance of journal bearings with oil grooves that were positioned at the maximum pressure location. They discovered that positioning the grooves at the maximum pressure location will cause 30-70 percent reduction in the load capacity of the bearing. However these results are not applicable to journal bearings with multiple axial grooves supplied with water from one end of the bearing only (Fig. 13.1). Theoretical calculations have also been undertaken to verify and augment the experimental data generated [16, 17], see Fig. 13.3.

The flow in a journal bearing supply groove has an important role in determining the performance of the journal bearing, represented by its load-carrying capacity and energy consumption. The flow in the grooves and land area determines the pressure field along the grooves and over the land area. Estimation of this pressure will give the load capacity (that can be supported by the lubricant film), coefficient of fluid friction and volume rate of flow. The data produced from measuring the lubricant flow is also used to determine the pressures at the groove


Pressure location



- 3971.1 N lineal

- 2976.4 N lineal

Pressure location


Fig. 13.3 Measured axial pressure distribution shown together with the linear variation in the groove (Hargreaves et al. [16])

Fig. 13.3 Measured axial pressure distribution shown together with the linear variation in the groove (Hargreaves et al. [16])

edges, which is an important consideration in the detail design of these bearings. It is important to note that there is no published literature on the type of flow phenomenon illustrated in Fig. 13.4, even though journal bearings have been used as stern-tube bearings in ships for many years. A much simpler bearing configuration has been studied experimentally by Salimonu et al. [20] and Burton [21]. In that work, the pressure profile across a blockage (rather than a groove) in a conventional bearing with shaft rotation and no axial pressure-induced flow along the groove was measured. It was evident that pressure jumps occurred near to and across the blockage (Fig. 13.3). This implies that such jumps will occur for the bearing being studied in the proposed project. This in turn emphasises the importance of the boundary conditions at the groove in any theoretical modelling used to predict bearing performance.

Some design procedures for axial groove journal bearings have been reported [22, 23] but once again, these do not cater for multiple axial grooves supplied with lubricant from one end only.

The configuration of the water lubricated bearing is similar to the submerged bearing studied by Pai and Majumdar [24]. The lubricant (water) is fed from one end through multiple axial grooves. Groove depth is kept at 3 mm but the number and size of grooves may vary. The lubricant film is generated in the non-grooved (land) region (see Fig. 13.1). The flow in this region may be both circumferential and axial. The lubricant flows out from the bearing ends axially [25]. They found that the maximum pressure in the clearance space of the bearing does not occur at the central plane but shifts closer to the outlet side of the bearing. This is because the lubricant is supplied axially. At the inlet, flow into the bearing takes place only in the unloaded region. At the outlet, flow takes place out of the bearing in the loaded region. The flow into and out of the bearing is maintained very much similar to that in a submerged bearing The pressure generated due to wedge action in the clearance space supports the applied load without metal-to-metal (or solid-to-solid) contact. Although the pressure in the grooved region has been shown in previous experiments to be constant, the pressure in the land region will be generated due to axial and circumferential flow components.

Majumdar et al. [26] studied the steady state and dynamic characteristics including whirl stability of water lubricated journal bearings having three axial grooves theoretically. The governing equation is the Reynolds equation in two dimensions which is solved for pressure numerically for the already defined boundary conditions. The Swift—Stieber boundary conditions are adopted in case cavitation occurs in the divergent portion. The pressure distribution in the clearance space gives the steady state characteristics in terms of load-carrying capacity, attitude angle, volume rate of flow and coefficient of fluid friction. These data help in design of the bearing. The fluid film bearings are prone to whirl instability to some extent; a further study was undertaken to ascertain the dynamic characteristics in terms of stiffness and damping of the fluid film and stability of the rotor bearing system. A first order perturbation method was used to find the dynamic pressures in the clearance space. The stiffness and the damping characteristics were calculated from the real and imaginary parts of the integrated pressures. These coefficients are then used in the equation of motion of a rigid rotor for estimating the mass parameter, a measure of stability. The already existing experimental data was verified by the theoretical solution. The pressure along the groove was found to be constant. It was seen that both load-carrying capacity and stability improve for the 18° angle grooved bearing, whereas friction variable is less. As expected, load capacity increases with the increase in eccentricity. Higher journal speeds leads to higher pressure development and results in a sharp fall in pressure at the bearing ends. High eccentricity will also give rise to high pressure development and will lead to higher flow rate. The coefficient of friction will increase with higher speed. This is due to the shearing action at high speeds. The mass parameter and whirl ratio are used as the measure of stability. The stability also increases sharply at a very high eccentricity ratio. It was concluded that bearing having smaller groove angle gives higher load capacity due to high pressure in the land region. The stability also improves when small groove angles are used.

Pai et al. [7, 8, 27] discussed theoretically the stability characteristics of water lubricated journal bearings having multiple axial grooves. A nonlinear analysis of a rigid rotor supported on journal bearings under unidirectional constant load, unidirectional periodic load and variable load was carried out. The time-dependent Reynolds equation in two dimension for incompressible fluid was solved numerically by the finite difference method, with Jakobsson-Floberg-Olsson boundary conditions, to obtain the hydrodynamic forces. Using these forces the equation of motion are solved by the fourth order Runge-Kutta method to predict the transient behaviour of the rotor. The analysis gives the orbital trajectory within the clearance circle. It was found that better pressure value distributions were obtained when JFO model was used as boundary conditions than using Reynolds boundary conditions. This helps in better prediction of bearing stability. The locus of the journal with a constant load has smaller excursions of the journal centre, when compared with a periodic load superimposed on the system. The trajectory for the variable rotating load was complex, but the journal locus lies within the clearance circle, indicating the bearing system can operate without any problem. Pai et al. [9] also studied the stability of four and six axial groove bearings using a nonlinear transient approach using the method described above.

Cabrera et al. [28] studied the film pressure distribution in water lubricated rubber journal bearings. The measurements of pressure indicated that the film pressure profiles are very different from those of the conventional rigid bearings. The relatively low pressures in the film caused significant rubber deflections but too low to produce viscosity changes. Integration of the pressure in the bearings showed that they operate in the regime of mixed lubrication. The behaviour of the bearings was theoretically investigated using CFD and compared with the experimental values conducted on the test rig designed (See Fig. 13.5).

The bearing test rig was used for 50 mm diameter bearings and length-to-diameter ratio of 2. The test shaft was made of chrome-plated stainless steel and supported by two hybrid bearings. The test loads were applied using a motorised unit, a de-coupling spring and a load cell. The loading unit allowed loads of up to 1,000 N to be applied. The tests were carried out with a shaft surface speed of 2 m/s and loads in the range 0-500 N. Production bearings manufactured by Silvertown UK Ltd. were used for this investigation.

The commercial CFD software, FIDAP, was selected as it allowed the solution of coupled structural and fluid problems. It has been found from experimental measurements and confirmed using CFD that complex film pressure distributions exist over the loaded staves. These distributions are a result of the interaction of the elastic bearing surface deflections with the film pressures. Peak pressures in the bearings are greatly reduced compared with conventional bearings.

Wojciech Litwin [29] has extensively studied the performance of water lubricated bearing through his experimental studies. The experimental setup is as shown in the Fig. 13.6. The diameter of the sea water resistant steel shaft was

Rig Layout
Fig. 13.5 Dynamic test rig layout (Cabrera et al. [28])
Test Litwin
Fig. 13.6 Stern-tube test rig (Litwin [29]). 1 Main shaft 0100 mm; 2 tested bearing bush; 3 baring sleeve; 4 covers with sealing; 5 static load leaver; 6 roller bearings; A pressure pickup sensor; B touch less distance sensors; C torqmeter

100 mm and the length being 700 mm. The shaft was connected with a clutch and torque meter to a electric engine so as to vary the rotating speed. The examined bearing is set in the steel sleeve and processed along with it. The sleeve is closed on both ends with the sealed covers which make it possible for the lubricating and cooling water flow inside it. This enables the bearing to work in different pressures

Lubrication Journal Bearing Mounted
Fig. 13.7 Measured pressures in water film in water lubricated bearing (Litwin [29])

up to the pressure of 0.6 MPa (Fig. 13.7). The load is exerted on the bearing in the static or dynamic way. The parameters which could be varied are as follows:

• Shaft rotation speed from 0 to 11 rev/s.

• The pressure of the water supplying the bearing, varying from 0 to 0.6 MPa.

These conditions are similar to the real small ship main shaft bearing. The pressure measuring (Fig. 13.6) system was a pressure sensor mounted inside the shaft and the signals were transmitted using a wireless system. The trajectory of the shaft was determined by having two pairs of non-contact transducers with an accuracy of 1 im on either side of the shaft (Fig. 13.8).

The second test rig called the Propeller Shaft bearing test stand is similar to the previous test rig except for the number of bearings to be tested are two. This test rig has the capability of measuring the fluid film pressure in the mid plane of the bearing.

Apart from the experimental tests the theoretical research was also conducted. The calculation results were obtained by using the author's original software based on hydrodynamic lubrication theory applicable to elastic bush material (EHL). The software is composed of two modules: the programme based on hydrodynamic lubrication theory, written in FORTRAN language, and the module based on the finite element method (FEM) in the form of macro to a commercial software. The calculations are performed in an iterative mode. A satisfactory result was obtained after twenty iterations (Fig. 13.9). Results of the calculations were verified experimentally. The author confirms that the applied calculation method is correct. In the author's opinion the calculation error does not exceed 20%.

Andersson et al. [30] have studied the load-carrying capability of water lubricated ceramic journal bearing. The possibilities of using advanced ceramics in water lubricated bearings were studied by performing tests on journal bearing lubricated with water. The materials studied were two aluminas (Al2O3) a zirconia toughened

Pipe With Multiple Bearings
Fig. 13.8 Stern-Tube test stand with two bearings [29]. 1 Main shaft 0100 mm. 2, 3 Stern-tube bearing. 4,5 Front bearing. 6, 7 Supports. 8 Covers with sealing. 9 Connecting pipe. 10 Radial load—discs. A Torqmeter. B Pressure pickup sensors. C Touch less distance sensors
Fig. 13.9 Calculated values of load-carrying capacity of a bearing in function of number of successive iteration steps [29]. 1st series—Stable solution was achieved. 2nd series—A convergent result was not achieved

alumina (ZTA), a partially stabilised zirconia (PSZ), a sintered silicon carbide (SiC). The tribological tests were performed using a journal bearing test rig built for this particular purpose. The ceramic test bearing was fixed into a sealed floating housing connected to a closed lubrication circuit. A shaft sleeve made of the same ceramic was attached to a steel shaft to be rotated. The test bearing housing was loaded by dead weights. During the test, the friction force and the temperature on the outside of the ceramic bearing were recorded. The tests were carried out with u 40 mm x u 30 mm x 40 mm shaft sleeves made of two aluminas (Al2O3), a ZTA, a partially stabilized zirconia (PSZ), a SiC, a reaction bonded silicon carbide (SiSiC)

and b'-sialon rotating in u 56 mm x u 40 mm x 20 mm journal bearing made out of same materials. The test was performed with a pair of new specimen and with de-ionised water as the lubricant. The water pressure was in the range of 3.5-6 bar and the feed rate 0.3-7 x 10-6 m3/s depending on the wear degree of the ceramic specimens. The temperature of the water before entering the bearings was 28 ± 5°C. Prior to testing all the specimens were ultrasonically cleaned and dried. The entire test was run at 0.4 m/s sliding velocity, which is equivalent to a rotational speed of 192 revolutions per minute. The load applied on the bearing was between 0.5 and 4.8 kN. It was found that tribological behaviour of the different ceramics in the present tests differed very strongly. It was found that silicon carbides performed well even at the highest loads applied. Both the SiC and the SiSiC materials experienced very low wear; their stable coefficient of friction was below 0.01 and their sliding surfaces became polished. Hence their load-carrying capability is at least 4.8 and 3.5 kN, respectively. The load-carrying capability of Al2O3 and ZTA ceramics was limited to 0.8-1.3 kN. The PSZ and sialon materials failed at the lowest load applied. To conclude we can summarise that evaluation of ceramics for possible use in sliding bearing assemblies lubricated with water and the best results were obtained with SiC and SiSiC, due to their ability to become tribochemically polished combined with their good thermal conductivity. These materials can be recommended for water lubricated bearing systems provided that the mechanical strength of the ceramics is sufficient for the application. In recent years, Yoshimoto et al. [31, 32] considered axial load capacity and stability of water lubricated hydrostatic conical bearings with spiral grooves.

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