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The purpose of this study is to investigate the influence of groove shape on the hydrodynamic characteristics of a journal bearing.

The computational fluid dynamics model also takes into account the cavitation phenomena and thermal effect, which can illustrate the lubrication performance of a journal bearing.

The hydrodynamic simulations of the journal bearing with the different groove shapes are conducted under different operation conditions.

Based on the numerical analysis, the suggestions are presented for groove shape selection and can be used to the design of a journal bearing under the extreme operation condition.

The purpose of this paper is to assess the turbulent thermohydrodynamic (THD) performance characteristics of an axially grooved finite journal bearing for a variety of simulated operating conditions.

The set of governing equations consisting the Navier‐Stokes, turbulent kinetic energy and its dissipation rate equations coupled with the energy equation in the lubricant flow and the heat conduction equation for the bush are solved to obtain the three dimensional steady state THD characteristics of journal bearings. The lubricant flow in turbulent regime is modelled using the AKN low‐Re k−ϵ turbulence model. The problem is formulated mathematically and solved numerically using the computational fluid dynamics (CFD) approach with appropriate boundary conditions.

It was found that shaft rotational speed has dramatic effects on the maximum temperature of the bearing and lubricant, also on the maximum hydrodynamic pressure and the oil flow rate. Besides, it was revealed that the clearance ratio and eccentricity ratio significantly change the performance of the journal bearing.

The paper presents a very useful numerical method for the prediction of the pressure and temperature fields inside the lubricant and thermal simulation of the bearing.

The paper provides the numerical simulation of the flow and heat transfer inside the journal bearing. Present computational approach is valuable to the practical modeling of the journal bearing operating under turbulent regime and in preparing the design charts in prediction of both hydrodynamic and thermal behaviors of journal bearings.

In a nine‐cylinder radial compression‐ignition engine, a single venturi passage 15 which serves for the admission of air and the exhaust of each cylinder is controlled by a valve 17 opened through a pushrod 19, and fuel is discharged through a nozzle, Fig. 8, controlled by a spring‐loaded valve 30 held off its seat by a stop 33. Fuel for all the cylinders is circulated through a manifold 40 and is delivered to the pumps through ports 41. Fuel is trapped by a plunger 39 and forced past a pair of non‐return valves 38 and a passage 34 to its appropriate fuel nozzle. The plunger is withdrawn by a spring 42, its forcing stroke being effected by a link 44 which bears on a camactuated lever 50, 51, 52. The stroke of the plunger is adjusted by toothed gearing 47, 48 which rotates a ring 46 carrying links 45 pivoted to the links 44. The fuel‐pump actuating cam 54 has four lobes and is driven through gearing 59 . . 62 at one‐eighth the speed of the crank shaft; this cam also operates the cylinder valves 17. The pumps may also be driven by a cam 73 having a single lobe and rotating at the speed of the crank shaft to which it is keyed. The high‐speed single‐lobe cam is operative when starting and times the fuel injection to occur just before top dead‐centre whereas the cam 54 times the injection so that it occurs 40 deg. to 20 deg. before the top dead‐centre. The starting gear comprises a motor 24 having a sliding dog clutch 25 which engages jaws 77 on a sleeve 76 mounted in a second sleeve 75. The sleeves are held against relative axial movement, but are each provided with diametrically opposed inclined slots 78, 79, Fig. 7, which impart to them a relative rotational movement. A pin 80 projects through the two slots into grooves 81 at the end of the crank shaft, and a spring 82 in the hollow portion of the crank shaft bears against a carrier 83 through which the pin 80 projects. During normal running, when the cam 54 actuates the plungers, the spring 82 forces the carrier 83 and the pin 80 to extend the sleeves 75, 76 to their rearmost position. The lobes on the cams 73, 54 are so arranged relatively that the one which is effective will shield the other from the levers 50, 51, 52. At starting, the jaw 25 is moved into engagement with the jaw 77 and the pin SO moves in its axial groove 81 to rotate the sleeves 75, 76 without imparting rotation to the crank shaft. The rotation of the sleeve 75 places the cam 54 in a retarded position so that its cam lobes are ineffective and pass under the lobe on the cam 73. As a consequence the speed of the fuel injection is increased, and as the engine starts the crank shaft overruns the jaw 25 and the spring 82 forces the sleeves 75, 76 and the pin SO to their rearmost position to establish normal running. The shifting of the cam 54 retards the actuating cam lobe 70 of the air inlet valve so that at starting the closing of the air inlet valve is delayed to allow part of the air charge to escape.

Herringbone groove thrust bearings are typically used in high-speed, light-load applications, such as spindle motors for hard disk drives. In the past researches, the effect of shaft misalignment was little considered. This study aims to reveal effects of shaft misalignment on the microscopic flow regime in the water-lubricated herringbone groove thrust bearing.

The liquid film in a thrust herringbone groove bearing was investigated by computational fluid dynamics. The effects of micro-grooves on the flow field were carefully explored. Two-dimensional liquid films at four different sites were examined for obtaining the rich flow field properties.

The distributions of pressure, temperature and water vapor volume fraction were obtained, the micro hydrodynamic effects were formed by the herringbone grooves and the effects of the shaft misalignment on lubrication and sealing performance could be found.

The influence of misalignment on the herringbone groove thrust bearing performance was investigated in detail. The obtained results could give the reference guideline for the bearing design.

Friction stir processing (FSP) is overviewed with the process variables, along with the thermal aspect of different metals.

With its inbuilt advantages, FSP is used to reduce the failure in the structural integrity of the body panels of automobiles, airplanes and lashing rails. FSP has excellent process ability and surface treatability with good corrosion resistance and high strength at elevated temperatures. Process parameters such as rotation speed of the tool, traverse speed, tool tilt angle, groove design, volume fraction and increase in number of tool passes should be considered for generating a processed and defect-free surface of the workpiece.

FSP process is used for modifying the surface by reinforcement of composites to improve the mechanical properties and results in the ultrafine grain refinement of microstructure. FSP uses the frictional heat and mechanical deformation for achieving the maximum performance using the low-cost tool; the production time is also very less.

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A method for optimizing net positive suction head required of axial‐flow pumps has been proposed by the present author, which is based on the two‐dimensional potential flow model and without considering the tip gap effect. The objective of the paper is to confirm if the method is just and feasible for the case of viscous fluid flow in impellers with tip gap.

A series of steady, three‐dimensional, noncavitating and cavitating, turbulent, incompressible flows of water through two axial‐flow pump impellers were calculated by using CFD code Fluent. The two impellers included a reference one with constant circulation at outlet and an optimized one with variable circulation designed with the author's method and code. In computations, the throttling and unthrottling approaches were used, respectively. Comparison of hydraulic performance, averaged flow variables at the impeller inlet and exit, flow in the tip gap, flow variables on blade surfaces and suction performance between the optimized and reference impellers was made.

It was confirmed that the optimized impeller has better hydraulic and suction performances. The method for optimizing with variable flow circulation profile along blade span at the outlet to impeller is proper and practical. Additionally, an unstable regime in the head curves of two impellers is presented. In the regime, a stall occurs on the pressure side of the blade and a hysteresis exists, which causes a hysteresis‐loop.

The effect of suction entry on flow is represented approximately by using a free‐vortex and uniform axial velocity. The diffusing component behind the impellers is not taken into account. The unsteadiness of flow is not considered, which would have a connection with stall pattern in an axial‐flow impeller.

The hydraulic and suction performances and flow variables of two axial‐flow pump impellers with tip clearance are obtained successfully with CFD. Stall and hysteresis as well as hysteresis‐loop in head curve are observed by using throttling and unthrottling approaches.

The stringent requirements for environmental protection have induced the extensive applications of water-lubricated journal bearings in marine propulsion. The nonlinear dynamic analysis of multiple grooved water-lubricated bearings (MGWJBs) has not been fully covered so far in the literature. This study aims to conduct the nonlinear dynamic analysis of the instability for MGWJBs.

An attenuation rate interpolation method is proposed for the determination of the critical instability speed. Based on a structured mesh movement algorithm, the transient hydrodynamic force model of MGWJBs is set up. Furthermore, the parameters’ analysis of nonlinear instability for MGWJBs is conducted. The minimum water film thickness, side leakage, friction torque and power loss of friction are fully analyzed.

With the increase of speed, the journal orbits come across the steady state equilibrium motion, sub-harmonic motion and limit circle motion successively. At the limit circle motion stage, the orbits are much larger than that of steady state equilibrium and sub-harmonic motion. The critical instability speed increases when the spiral angle decreases or the groove angle increases. The minimum water film thickness peak is at the rotor speed of 4,000 r/min for the MGWJB with Sa = 0°. As rotor speed increases, the side leakage decreases slightly while the friction torque and the power loss of friction increase gradually.

Present research provides a beneficial reference for the dynamic mechanism analysis and design of MGWJBs.

IT is well known that war gives a great impetus to development in many fields, not least of which is that of aircraft propulsion. Such was the case in World War II, when great strides were made, but it is interesting to note that the pace has hardly slackened in the years following its conclusion. This is perhaps because of the ‘cold’ war which took its place, or perhaps because the introduction of jet propulsion has stimulated thought and action in realms beyond the dreams of the piston engine era. Whatever the cause, the results are apparent and this is a suitable moment to look back and measure the progress of the past seven or eight years.

This paper gives a review of the finite element techniques (FE) applied in the analysis and design of machine elements; bolts and screws, belts and chains, springs and dampers, brakes, gears, bearings, gaskets and seals are handled. The range of applications of finite elements on these subjects is extremely wide and cannot be presented in a single paper; therefore the aim of this paper is to give FE researchers/users only an encyclopaedic view of the different possibilities that exist today in the various fields mentioned above. An Appendix included at the end of the paper presents a bibliography on finite element applications in the analysis/design of machine elements for 1977‐1997.

MAKERS OF STATIONARY GAS TURBINES for industrial uses claim that the lubricating cost of medium‐size power units ranges between 1 and 2% of the fuel costs. Comparative figures for average Diesel engines are 5 to 10%. These savings have an effect on the total running cost economy. Detailed oil consumption figures from industrial gas turbine operators have not yet been disclosed and for similar reasons, it is not possible in this survey to discuss individual design features of all units which comprise the various lubricating systems.