The authors investigate the characteristics of torsional vibration displacement and stress of a crankshaft with a viscous fluid damper from the experimental and analytical viewpoints. The obtained results in this paper are as follows: 1) The phase angle between the inertia ring and the case of the viscous fluid damper is considerably large, so the phase angle on the mode of torsional vibration cannot be neglected in this kind of the vibration systems. 2) The engine speed of the maximum torsional vibration stress at one point of the crankshaft is different from that at the other point of the crankshaft. 3) The engine speed of the maximum torsional vibration displacement does not necessarily correspond with that of the maximum torsional vibration stress. 4) The torsional vibration displacement and stress in a crankshaft with a viscous fluid damper can be approximately calculated by using the 3 dimensional analysis of forced vibration by the transfer matrix method, in which the values of the complex viscous damping coefficient obtained with the experiment of a vibro-viscometer are adopted.
The torsional vibration damping of diesel engines is assumed to consist of viscous friction damping and hysteresis damping. Their damping ratios are divided into two kinds of non-dimensional functions, respectively; that is φ1 (cr, D), φ1 (dc, D) composed mainly of engine size, and φ2 (n, p), φ2 (n, p) composed of natural frequency and number of equivalent masses. The characteristics of engine damping have been discussed in this paper by investigating the tendency of these nondimensional functions. The obtained results are as follows: 1) The torsional vibration damping of an engine ζecan be expressed by the following equation, ζe=ζr+ζh=φ1 (cr, D) ⋅φ2 (n, p) +φ1 (dcD) ⋅φ2 (n, p) where, ζr =φ1 (cr, D) . φ2 (n, p) ; damping ratio due to viscous friction damping, ζh=φ1 (dc, D) . φ2 (n, p) ; damping ratio due to hysteresis damping, cr =viscous damping coefficient for one cylinder, dc= mean diameter of crankshaft, D= cylinder bore, n= number of equivalent masses, p= frequency ratio. 2) It is difficult to estimate accurately the value of the non-dimensional equivalent damping function φ1 (cr, D), because any equation cannot calculate the value of crwith high precision. The value of φ2 (n, p) is greatly dependent on natural frequency. 3) As the formulae of Lewis has been adopted to estimate the hysteresis damping energy, the damping ratio due to hysteresis damping ζhis proportional to θ0.310 (where, θ10; angular displacement) . The value ofφ1 (dc, D) varies from 5×10-3to 10×10-3, independent of engine size. The value of φ2 (n, p) for all torsional vibration is less than unity. Then, ζhhas values of the order of 10-3.
The loss coefficient of valve is the important factor on design of piping system, and various test results were reported in the past. However, the coefficient depends on the construction, type and size of valve. The reports by The Ship-Machinery Manufacturers' Association of Japan on 1960 and 1961 clarified the loss coefficients of JIS F marine globe valves, angle valves and sluice valves of small size up to 130 mm bore. In this paper, test results of the loss coefficients of currently used 200 mm bore globe valves and sluice valves complied with JIS F for marine use and JIS B for land use, and butterfly valve for city water service are reported, and relation of coefficient of various valves are discussed.
The paper describes results and evaluations of investigation concerning emulsified fuel oil combustion on marine diesel engine. The long-term trial by using emulsified low grade bunkers has been carried out on ocean-going diesel vessels by Showa Line LTD. According to the trial, the fuel oil saving has been attained and also various other merits such as lesser fouling of exhaust economizer have been obtained.
Along with recently distinguished clamour for improved performance and reliability of the gas and steam turbine now in wide use both for marine and land purpose, the renovated design techniques have become indispensable for the rotating blade of the turbine in the sphere of fluid hydraulics and vibration stiffness. In reality, the rotating blade is used where the centrifugal force works, besides being complicated in its style in a form of curvilinear unit, furthermore, at the time of its designing, extremely accurate calculation of estimated vibration is required for preventing inevitable resonance due to harmonic vibration. A practical analysis method must be therefore be established, taking these specific characteristics of the blade into consideration. Accordingly, pinpoint accurate calculation will naturally necessitate large sized FEM models, with simplified disposal of Input/Output data and reduced cost of calculation becoming pivotal requirements for practical use. In order to satisfy these prerequisites, this report was prepared to illustrate in summary as to how we have developed the practical FEM design/calculation method and then put into physical utilization at an improved accuracy, as we have confirmed from our shop tests for comparison.
With respect to analysing coupled torsional-axial-flexural vibrations of the crankshaft, a simplified crankshaft model of single throw made of uniform steel wire of a small diameter is introduced, and its influence numbers for all kinds of loadings conceivable are derived and shown in two tables. Though they are only qualitatively true for actual shafts, the method of deriving influence numbers is commonly applicable in principle.