Methods of electric arc welding with metallic electro'es now used are classified into three; namely, hand welding, automatic welding and semi-automatic. However, in these methods the electrode has to be fed and moved along the weld line as it is melted.
The new method introduced by the author provides for the laying of the electrode directly (in the case of a bare electrode some clearance between the rod and the base metal is required for laying several) short pieces of asbestos yarn transversely between them) on the weld line of the base metal, and as soon as the arc is started by shorting the electric current between the tip of the electrode and the weld metal with a pieceof wire, the wire is removed, then the arc will move continuously at a constant speedand a constant arc length from the tip of the end (which is connected with the electric terminal wire), and the welding is completed.
The new method can be used in vertical, horizontal and overhead welding without any difficulty, just as easily as downward welding.
The new method requires neither the personal effort of the welder, as in hand welding, nor does it restrict the size and capacity of the welding machine, as in automatic welding.
The electric welding machine and electrode now in use can be employed with the new welding method without any alteration to them whatever.

抄録全体を表示

A hydraulic testing machine for 3, 000 ton compression and 1, 500 ton tension tests was installed at the Kure Naval Dock Yard, in September 1934.
It is designed and constructed with the latest improvement to suit the testing of ship's structure, the same line with that installed in the “Material-Prüfungsamt” in Berlin-Dahlem.
The principal dimensions of the machine are as follow:-
Overall length of the machine when the tie spindles are moved outward. 28, 000mm
Extreme width of the machine. 4, 400"
Available stroke of ram at compression test. up to 400"
Available stroke of ram at tension test. up to 500"
Maximum length of compression test piece. 15, 100"
Maximum width of compression test piece. 2, 000"
Maximum length of tension test piece. 12, 100"
Maximum width of tension test piece. 2, 000"
Accuracy of the machine, denoting error 3% at 150 tons, is 1% for 300 tons and decreases to 0.15% at the maximum load for tension test.
The authors give a description about its mechanism and operation, and also on explaining the results of tests for riveted and welded joints, they point out the indispensable value for the model testing of complicated construction.

抄録全体を表示

The steam conditions for marine engine plants using marine Scotch boilers are usually limited, in our country, to the steam pressure of 16kg/cm² G., steam temperature of 285°C and the vacuum at the condenser top of 710mm. Now for this vessel we preferred them as 17.5kg/cm² G., 330°C and 717mm respectively, and made some improvements in the design of machinery parts. For the main turbine, Parsons' coefficient has reached 2, 125m²/sec².kcal, which is quite an improvement compared with the former usual value of 1, 500-1, 800.
Trial trips were made on the 9th and 13th August for the first of 3 sister ships, The “Heito Maru”, on the 17th and 19th September for the second ship “Taito Maru”and on the 31st October and 2nd November for the third ship “Shoka Maru”. The results are tabulated in the attached table. The coal consumptions for all purposes were 0.466, 0.462 and 0.492kg./s.h.p./hr., respectively, and the whole plant efficiencies showed more than 17%. Particulars of machinery parts are shown in the attached table, and the general arrangement in Fig. 1.
Propelling machinery consists of one set of Mitsubishi-Zoelly double reduction double helical nodal drive geared, cross compound, all impulse turbines of the latest improved type. For reduction gear, taking special consideration for smooth running and for elimination of troubles under undue stresses due to the torsional vibration of shafting and also the end play in each pinion, the second pinion shaft is flexibly driven by a long nickel steel shaft provided with a special flexible coupling between the first wheel and the flexible driving shaft. All discs of H.P. ahead and astern turbines are solid with the rotor shaft, and those of L. P. turbine, ahead and astern, are shrunk on the rotor shaft with keys.
The main condenser is of Mitsubishi-Contraflo type and is located on the port side of the main turbine.
Three boilers of single ended, multitubular marine type, each 4, 600mm diameter and 3, 650mm length with a total heating surface of about 232.9sq.m. and a total grate area of about 17.5sq.m. are constructed under the special survey of Teishinsho for a working pressure of 17.5kg/cm², Fig. 2 shows the sectional arrangement. The superheaters are of a combination of smoke tube type and Mitsubishi combustion chamber type. The saturated steam is superheated primarily in the former and secondarily in the latter, up to the final steam temperature of 330°C. The construction is shown in Fig. 3. Regulating and by-path arrangements are provided, so that any required proportion of mixing saturated steam with superheated steam is easily obtained. Superheaters will be thrown out of connection quite easily whenever required. These connection diagrams are shown in Fig. 4.

抄録全体を表示

A new design of an oil tanker is proposed, in which the special attention is paid to the following points:-
(a) To reduce the abrupt change of strength or form as far as possible.
(b) To make the longitudinal members continuous.
(c) To fit the necessary girders in the holds.
To solve these points, the following means are taken:-
(a-1) Boundary flanged pieces are used in lieu of boundary angles, connecting the transverse bulkheads to the longitudinal members, such as, longitudinal bulkhead, shell and deck plating, which will reduce the abrupt change of forms at these connections.
(a-2) The upper deck corners on ship sides are made round, just like the turn of bilge, which will also reduce the abrupt change of form to a remarkable extent, and, as a result, the strong gunnel bars will be omitted, and the sheerstrakes and stringer plates are lightened.
(a-3) In the construction of the bridge, the girder works are accepted instead of superstructure constructions, without shell and end bulkhead plating, which will eliminate the abrupt change of strength and form caused by the superstructure of the bridge deck, and make the water on deck free from there very rapidly. But it will be necessary to prepare a break water bulkhead of simple construction in front of the poop front bulkhead, against the rushing water from the ship bow toward the poop front bulkhead.
(a-4) Longitudinal members, such as, deck plates, longitudinal bulkhead plates and their connecting angles are arranged to co-operate in good balance.
(a-5) The landing edges of the lap joints of the upper deck plates at their seams are not scarphed, but taper liners are used to make sound joints.
(a-6) The oil hatchways of the upper deck are made in circular form instead of therectangular ones, and the beams continuous across the hatchway.
(b-1) The stringers and keelsons are made continuous through the transverse bulkheads, using watertight flanged collar pieces, which will eliminate the unsupported areas of the shell plates at the bracket connections of these members.
(b-2) The longitudinal members, such as, deck girders, side stringers, etc., are not slotted, but the beams or frames are cut off at these longitudinals and bracketted on them.
(c) As the clearness of hold is not needed in oil tanker, necessary girders are fitted in the holds longitudinally, transversely, diagonally, and also pillared. Both the end bulkheads of tanks specified, being connected by the boundary flanged collar pieces to the deck and shell, and stayed by the longitudinal girders, the construction of such tanks is just like that of the boiler, which is, of course, the most economical, stable and efficient construction.
(d) The flanged pieces are made as flanging plates, or rolling out from the ingot like section bars rolled. The workmanship of the flanged pieces, such as, joggling, bending or corner piece making, is much easier than that of angles or other section bars.

抄録全体を表示

For the section of a hollow-cylinder in heading, we can apply the torsion stress function in the following form:-
Ψ=k(1/aⁿ-1/γⁿ•cosnθ)-γ²+a²
where n=0 or integer, a=radius of a circle, k_??_0.
The function holds good in the outer region of a circle, and the inner periphery can be obtained by solving the following equation with respect to a parameter γ, .
k(1/aⁿ-1/γⁿcos nθ)-γ²+a²=0.
The outer periphery can be determined by taking
Ψ=C₂=a²-b²+k(1/aⁿ-1/bⁿ)
where b>a.
Then we have the stresses in two directions
Zθ=tG(γ-1/2knγ^-1-ncosnθ),
Zγ=tG(1/2knγ^-1-nsinnθ).
Between two stresses, it is necessary to consider the value of Zθ on the outer boundary, which becomes maximum at the angles θ=(2s-1)π/n, where s=1, 2, 3……n

抄録全体を表示

On the existing ships of about from 250 to 385 feet in length, usually there are two cargo hatchways for each of their fore and aft holds irrespective if there is a bulkhead in the hold or not.
It is not only convenient for ships to have one cargo hatchway to each hold to handle the cargoes of the usual size, but it is quite necessary to divide each of fore and aft holds by a bulkhead from the safety point of view, though their permeability, as shown in Table 1, is less than 60%.
All of the above stated ships are able to have a bulkhead in each of their fore and aft holds without disturbing the arrangement of their cargo hatchways.
On the other hand, though an ideal arrangement of the cargo handling gear is that the time to be needed for handling cargo is made equal for all holds. In fact, as shown in Table 4, it is difficult to do so for several reasons.
Thence the reasonable position of the bulkhead which divides the hold is such that the adjacent holds will have the same permeability.
The position of the bulkhead to be fitted in the hold is not equal for all ships as shown in Fig. 1.
Therefore, at the early stage of the design, it is convenient to fix the hold length by the following formula :-
l3-L3=(l2-l1){3/2(100/μ-1)-3/2(100/M-1)}
μwhich was derived from the well known formula
l3=(l2-l1)×3/2(100/μ-1)+l1
where l1, l2 and l3=permissible lengths of the referred ship,
μμ =permeability of the referred ship,
L3 =permissible length of the new ship,
M =permeability of the new ship.
The results calculated are shown in Tables 2 and 3.

抄録全体を表示

The belief, that an unbalanced force applied at the loop of the deflection curve in natural vibrations of a ship or an unbalanced moment of force applied at its node excites large vibrations under resonance, seems to have been prevalent among naval architects and engineers until recently. It appears that Schlick* was the first to give this conclusion, and the same view, being apparently established, was followed up even by some new writers such as Lewis** and Abell §.
Although I stated in a previous publication §§ that such a fact is not probable to exist from a dynamical point of view, the problem was not perfectly studied owing to certain difficulties. The present closer examination of the subject however confirms that the conception which has dominated the opinion of people is rather contradictory to nature with respect to the vibrations of a ship.
With a view to investigating the problem in the simplest possible way, I took a free-free uniform bar and calculated three cases; namely, (i) its flexural vibrations due to an unbalanced force (Appendix A), (ii) its torsional vibrations due to an unbalanced torque (Appendix B), (iii) its torsional vibrations due to an unbalanced torque moment(Appendix C). Although, in spite of its importance, the case of flexural vibrations due to an unbalanced moment of force was not studied at this juncture through some reason, it is nevertheless obvious that the nature of the problem is quite similes to the one in (iii) for torsional vibrations, and its full explanation will appear in the near future, as it is associated with some other problem.
The result of the present problem shows that, under resonance, the amplitudes of the ship's vibrations in the very vicinity of the engine room, depending on the magnitude as well as the frequency of the unbalance, either of force or of moment, besides the stiffness and the mass of the engine, are however the same wherever that engine room may be located. We also come to a conclusion that, in the case of the unbalanced force, the amplitude of a specified point in a ship in resonating forced vibrations with a given period differs in accordance with the difference in the position of the engine room, namely it is inversely proportional to such a particular amplitude that the point corresponding to the position of the engine room should have if the ship were to vibrate undamped and freely with the period under consideration Thus, should the engine room be at the node of free vibrations of the ship, the amplitudes of forced vibrations would be infinitely large, at any rate, in an idealized case, while, if the engine room were at the loop of the same free vibrations, the amplitudes of the forced ones would assume minimum values. The reverse is the case provi ed the unbalanced moment be applied in lieu of the unbalanced force.
Although accurate information is scanty now to me with respect to the details of unbalances in engines and in their accessories to cause sensible vibrations to ship structures, it seems however possible for some auxiliary engines to be more or less influential origins of the vibrations. Were such engines having unbalanced force set at a node of the principal free vibration of the ship, as will probably happen relatively frequently, excessively large vibrations of the ship's structure, particularly at the midship part as well as at its ends, would be expected.

抄録全体を表示

When an ordinary Merchant ship, afloat on deep still water under no wind condition, is proceeding ahead by her own inertia only, her propellers having been completely stopped and her speed length ratio K/√<L> (where K is speed in knots and L is L. P. P. in feet) is less than 0.4, resistances acting upon the ship are principally skin resistance and the resistance of stopped propellers dragging through the water, as the wave making resistance is practically negligible. The present paper deals with the above mentioned dragging resistance of stopped propellers of a ship longer than 400 feet, proceeding at a speed from 8 to 3 knots.
The loss of kinetic energy in ft. tons of a ship of W tons displacement, in the course of an interval between the instant when her velocity was v₁ ft. pre sec. and the instantwhen her velocity became v₂ ft. per sec., can be calculated as follow:-

抄録全体を表示