The vibration characteristics of rotating shaft system should be computed considering the dynamic stiffness of bearing foundation, especially if this stiffness is rather low. But the accurate characteristics of such a system has not been computed, as far as known, so that the dynamic stiffness of the bearing foundation can have hardly been obtained.
In this paper, a method to estimate the dynamic stiffness of bearing foundations by measuring their compliances dynamically is presented. On the other hand, the critical speeds and damping ratios of a model shafting system, consists of shaft, bearings and their foundations, was measured on the testing rig, and was compared to the estimated ones by using the dynamic stiffness obtained by the above mentioned method.
There is good agreement between the measured and estimated ones, though the stiffness of bearing foundations of the testing rig is relatively low compared with actual machinaries. It is expected that the method described in this paper is useful to predict the vibration characteristics of rotating shaft system of actual machineries.

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In zweitem Bericht wurden die Meßergebnisse vom Charakter des Viskose-Drehschwingungs-dämpfer beim schnellaufenden Dieselmotor betrachtet. Im diesen Bericht werden die Drehschwingungsmessungen an einem langsamlaufenden Dieselmotor gemacht, der mit den Viskose-Drehsthwingungsdämpfern der mäßigen Größe ausgerüstet ist, wenn man die Viskosität des Wirköles im Dämpfer vielerlei wechselte, und daraus wird der Charakter des Viskose-Drehschwingungsdämpfers beim langsamlaufenden Dieselmotor ermittelt. Der Schluil lautet wie folgend:
1) Die Meßergebnisse vom Schwingviskometer, d.h. die Werte von tan δ' in den Abb. 8-11 des ersten Berichtes sind auf den Original-Viskosedrehschwingungsdämpfern beim langsamlaufenden Dieselmotor mit den Resonanzfrequenzen im Gebiet j=26-30 Hz beinahe übertragbar.
2) Wenn man auch für beide schnell-und langsamlaufenden Dieselmotor den optimalen Dämpfungsbeding des Dämpfers mit χ²+α²=1, 0 oder χ+α=1, 0 irgend denkt, dann hat dieser Unterschied jedenfalls in der Tat kein Problem. Hier ist χ dimensionslose Dämpfungswiederstands-konstante und α dimensionslose Federkonstante. Aber aus den früheren und diesen Experimenten paßt es sich besser, daß man im allgemeinen für die Viskose-Drehschwingungsdämpfer bei Mehrzylindermotoren den optimalen Dämpfungsbeding mit χ²+α²=1, 0 denkt.
3) Das Verhältnis zwischen χ und α erhiert man wie folgend.
a) Für den schnellaufenden Motor (Resonanzfrequenzen f_n=160-190Hz) ; α=3, 45χ^3.6.
b) Für den langsamlaufenden Motor (Resonanzfrequenzen f_n=26-30Hz) ; α=0, 2χ^1.7.
4) Wenn die Dämpfermasse an beide schnell-oder langsamlaufenden Dieselmotor nicht zu klein ist, hat die Motordämpfung auf die Gräße der Dämpfungsamplitude keinen Einfluß.

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Various experimental equations (designated as experimental equations hereinafter) have hitherto been applied on torsional stiffness calculations of crank shaft in the form of equivalent length. However, all of these experimental equations had been obtained more than 50 years ago. Recent engines become substantially different from engines of those days in respect of weight, speed and so on, and characteristics of crank shaft are also conspicuously changed. Accordingly, experimental equations produce large errors and have little universality for diversification of crank shaft shapes. Therefore, the equivalent length is now obtained from actually operated engine data or estimations from an engine having similar characteristics, thus an experimental equation being used with its correction factors properly determined by an engineer in charge.
In this study, first, torsional stiffness of crank shaft was dynamically measured by the use of a vibrator, and good coincidence between the measured results and values calculated back from measured results of torsional vibration in an actual engine, was made clear. Then, torsional stiffnesses were measured with the shape of the crank shaft changed variously, and simultaneously torsional stiffnesses were obtained by the FEM, thus calculation accuracies of more than and inclusive 96% being obtained with respect to actually measured results for all shapes of crank shafts.

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On board a turbine ship equipped with a ducted propeller going ahead at a diminishing speed following crash stop astern operation, big noise occurred at its final reduction gear box. It was caused by collision of the bull gear with the final pinions of the high pressure turbine shafting, while no one had clarified the reason why they so randomly as well as violently collide at the speed range. The author describes some of the given data and his reasoning in the past.

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In recent years, two-stroke slow speed diesel engines with long stroke, large bore and a small number of cylinders are on the increase, to get higher propulsion efficiency of the engine and the propeller and less maintenance of the engine.
Diesel engines with a small number of cylinders, however, have some kinds of vibration problem. Torsional vibration problem is one of them. In case of five-cylinder engines, for instance, 1-node 5th-order torsional vibration is often dangerous for normal shafting design, so some countermeasure is necessary. In such cases, ‘the tuning disk with hydrodynamic damping’ is useful.
Of course, to get the optimum tuning for each different shafting, computer aided design (C.A.D.) is very useful.
Using Geislinger damper, optimum tuning can be obtained by choice of suitable elasticity and damping.

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Isolation properties of antivibration mounts were tested with models of the machinery casing and the ship-like structure. The mobilities of the machinery base plate and foundation bed plate of those models were found to be approximated to rigid masses at low frequencies and infinite plates for higher frequencies compared with the first local plate resonance.
Using these results and regarding resilient mount as massless spring, isolation properties of the soft and hard resilient mounts and the reduction method of the input vibration power to the ship structure were investigated.
In addition, a new type of isolation system in which a damping plate was inserted between the upper and the lower resilient mounts was devised and its superior vibration isolation effect was verified experimentally with the above mentioned models.

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