In the previous report, a theoretical solution, in which the numerical calculation was comparatively simple due to the rapid convergency of the infinite series, was given for the flow through an arbitrary lattice, and especially the effects of the lattice arrangement and profile form were numerically clarified. In this report further researches are carried out on the inclination of the lift coefficient curve, the moment coefficient and the surface pressure distribution. These characteristics are illustrated on Clark. Y., ogival and circular are profiles arranged in various lattice conditions.

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In author's previous paper theoretical investigation was carried out on the mutual interference between the impeller and the guide vane of axial flow pumps or fans, on the assumption that the effects of the interference were approximately similar to those of two equivalent vortex sheets. In this paper the more accurate theory is stated in which one of the impeller of the guide vane is replaced by an exact lattice of hydro- or aero-foils, and the effect of the other is treated equivalently as that of a street of vortex sheets, although this method cannot be applied to the case when the blade number of the impeller is unequal to that of the guide vane. An numerical example is shown for comparison with the results by the previous approximate method.

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A method is stated to calculate theoretically the mutual interference between the impeller and the guide vane, assuming that the flow is approximately similar to the potential flow through the two equivalent street of vortex sheets-one of them is moving and the other resting. It is clarified that the lift acting upon the blades of the impeller and the guide vane varies periodically due to the interference, and that the amplitude of variation increases rapidly as the gap between the impeller and the guide vane becomes smaller. Numerical calculations are also carried out for comparison with the experimental results obtained by Prof. C. Pfleiderer.

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In this paper, a method of calculating airfoil section form with given pressure distribution is treated considering the effect of viscosity. The velocity on the airfoil surface and the magnitude of circulation around airfoil differ between perfect and viscous fluids and these differences arise from the formation of boundary layer on the airfoil surface. In this method, the boundary layer thickness is calculated theoretically at first, then the above-mentioned differences are estimated assuming the thin layer covering airfoil with height of displacement thickness of the boundary layer is filled with streaming fluid issued from the imaginary source distribution on the airfoil surface, and hence corresponding pressure distribution of the perfect fluid is obtained. Then airfoil form is calculated for this corrected pressure distribution by the method already obtained by the senior author in the case of perfect fluid. Both cases of isolated airfoil and wing lattice are treated. An example of isolated airfoil (N.A.C.A. 4412, α=12°, R=3·10⁶) is shown in the original paper.

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In the previous papers, the author stated that the compressibility and lateral component velocity can be neglected for liquid pulsating flow. Now the author studied theoretically the pulsating flow in circular pipe of incompressible fluid, and obtained the instantaneous velocity, discharge and fluid resistance.

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The motion of a incompressible fluid along a vibrating wall is analysed by Havelock, but it is concerned with the vibration parallel to the wall surface, and the case the vibration is normal to the surface gives a more interesting problem. But in the latter case, in relation with the wave motion the compressibility of fluid must be taken into account. In this paper, along a plane wall a viscous laminar flow of a compressible fluid with uniform velocity gradient is assumed, and the influence of lateral vibration of wall, especially to the frictional resistance is analysed. As a conclusion, the ratio of increase of resistance is propotional to the square of Mach Number of vibration velocity of wall.

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In the last report, the sealing performance of a screw-type viscous pump with a screw groove of the rectangular cross-section was treated theoretically. For the purpose of determining the most suitable shape of the screw groove for the sealing performance, the author examine in this paper theoretically a rectangular, semi-circular, fan-shaped and triangular cross-sections respectively. As the results it is seen that in order to obtain the greatest sealing action i.e. the maximum pump lift without delivery, we should select the following shape of the groove ; a) The depth of each screw groove be as shallow as possible. b) The triangular section is the most suitable when the ratio of the depth to the width of the groove remains constant. c) Among the triangular sections it seems that the isosceles one is superior to the others.

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In the 2nd and 3rd reports the author presented already some theoretical conclusions on the sealing performance of a viscous pump. To certify those conclusions experimentally, he prepared a set of testing apparatus and 20 rotors with various shapas of the screw groove. Consequently such experimental results show a good agreement with the previous theory, namely a) The pump lift is proportional to N, L and ν, and inversely to δ〓 respectively. b) The theoretical results for the coefficient of sealing performance Sis verified experimentally. c) The theoretical comparison of various shapes of the groove is also ascertained to be correct by these experiments. d) So far as the sealing action is concerned, the multi-threaded rotors behave in the same manner as a single-threaded one. e) All theoretical results are confirmed experimentally not only regarding a "shaft" screw-type viscous pump, but also a "cylinder" one.

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Ship's speed-meter which has been designed by the authors, is based on the following principle. The force which produced by the dynamic pressure of Pitot-tube, is always balanced with the magnetic force by the automatic control of the magnet excite current. Then the instantaneous velocity of ship is given by the excite current and the integrating value of the excite current indicates the distance. By this ship's speed-meter the authors completed the experiment in the laboratory and ship.

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The Technical viscometers, which include Redwood, Saybolt, and Englar viscometers, have various defects. Therefore, the author carried out this experiment to find the characteristics of viscous flow through various kind of tubes and orifices into the air vertically, and to make a new Technical absolute viscometer. More than forty numbers of tubes, short tubes and orifices (the diameters of which being 0.03-0.53 cm. and the ratio of the length of the tubes or orifices to their diameters 0.024-124.6) were taken for this experiment. This paper presents the apparatus, the method of experiment, the flow pattern from the end of the tube or orifice into the air, and the relation between the coefficient of discharge and the Raynold's number.

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This paper presents mainly the various losses of head for viscous flow through the tube and orifice. Namely, the loss at the entrance, the loss at the inlet-length, and the loss in out let flow, etc. The coefficient of the additional-loss of head Δξ=H/(v²/2g)-(64.0/R_e) (l/d) has been obtaind in the form of the function of l/d and R_e, and when the tube length is longer than the inlet-length, the additional-loss of head Δξ_∞ is denoted by the function of R_e only, where R_e is the Raynold's number dυ/ν, H the total head cm, υ the mean discharge velocity cm/sec, v the kinematic viscocity cm²/sec, and l/d is the ration of the length of the tubes and orifices to their diameters.

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The general equation to determine the form of a notch is obtained in the following form, [numerical formula] when the theoretical discharge Q is given by a function of the head of the form φ (h), where h is the head over the notch and x=f(y) is the width of the notch at the height of y from the crest. And also above result is applied to several types of notches.

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The manometer, attached to the instrument for artificial pneumothorax, is used for measuring the breathing pressure in a thorax cavity. On that occasion, the pneumothorax needle forms a neck of air flow and the manometer has its own property. So it is inevitable that the indication of manometer surface is attended with consider able error and time lag. Author made some experiments on the relation between pressure variation and manometric indication by means of photographic self-recorder, and made the comparative study of the experimental results and the theoretical ones.

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With regards to the performances of jet-apparatus such as "ejector", or "injector", so far as the authors sware, it seems that there is any conclusive results neither theoritecal nor experimental. In this research series, they have intended to investigate phenomena of flow in the jet-apparatus, and then to obtain reliable data applicable to designing a high efficiency one. This report presents a new method to compare the efficiencies of jet-apparatuses. In the begining of this paper, the efficiencies of many kinds of jet-apparatuses practically used as locomotive blast were compared by this method. Then the effect of changing the length of mixing chamber on the efficiencies was investigated for the most basic type of circular section, and the length of the mixing chamber to give the maximum efficiency was also obtained.

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In this report, the behaviour of the flow in the mixing chamber was investigated. Using a new designed small pitot-tube, the velocity and pressure at every point in the mixing chamber were measured. As the results, the behaviour of the mixing process was made clear to some extent ; i.e. the distribution of the axial velocity v_a and the pressure p_s in each section were obtained ; the distribution of the radial velocity v_r was calculated basing on a new idea from the results. Thus, the mode of flow in the mixing chamber was shown by velocity-vector and streamline diagram. The distribution of the energy per unit area per unit time, and the total energy through each section per unit time, as well as the efficiencies and the losses of the apparatus have been calculated and graphically represented.

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Author treated upon the correlation between the oil pressure produced in the surbo-cylinder and the pressure and viscosity of jet oil in jet-pipe principle automatic controller. In addition we treated upon the correlation between the operating force and speed of the surbopiston and the removal of the jet pipe. Author experimented above correlation by using the apparutus showed in Fig. 2.

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With regard to the turbulent flow in a smooth pipe, assuming the following formula, and deciding the constant k adopting the experimental values made by Nikuradse, the author gives fairly accurate experimental equations for thier simple experssion. [numerical formula] l=mixing length, r=radius of pipe, η=y/r, y=distance from pipe center, V=velocity at pipe center, u=velocity at y, [numerical formula], γ₀=shear stress at pipe wall, p=density of fluid.

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Uniform spiral flow in a central exhaust pipe and jet spiral flow which blows tangentially into one end opened cylinder from one direction, are fundamental ones in a cyclone dust collector, and so they have been measured by a cylindrical pitot tube, and discussed. The spiral angles of these flows are not affected by flow quantity, and the velocities are proportional to flow quantity. In the case of jet spiral flow, when the cylinder is extended to a certain limit, its pressures rise and its spiral angles decrease suddenly. The axial velocity distributions are not radially constant, but decrease inwards almost straightly as well as spiral angles.

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We could not make clear the hydraulic loss of the accelerated or retarded flow which is caused by openning of the distributing valve. In this paper, the author studied, by means of operational calculus, the retarded flow vhich changed from the steady flow. The result showed that the same treatment for the steady flow can be applied when the kinematic viscosity of the flowing liquid is large and the pipe is narrow.

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In this report next three cases were solved under the assumptions that the flow is laminar and the compressibility effects of the fluid can be neglected. (1) When the pressure gradient given to the pipe begins to increase or decrease in proportion to the time. (2) When the pressure gradient changes instantaneously. (3) When the pressure gradient begins damped oscillations. (including exponential damping and harmonic oscillations as special cases). For each case the velocity distribution in the pipe and the discharge per unit time were found, and the transient state of the flow was also explained. From these three examples mentioned above, it was shown that the flow differs widely from that of Poiseuille when non-steady pressure gradients were given along the axis of the pepe

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In this paper, the velocity distribution in the laminar boundary layer of the incompressible fluid is treated by the variational method. The functional J (u) is obtained, which has stationary value under a certain additional condition, when the velocity distribution u satisfies Navier-Stokes' equation and the equation of continuity. To solve the problem of the calculus of variation thus constructed, Ritz's method is used. By the above mentioned method the difficulty of solving non-linear differential equation is avoided. For the first time, this method is applied to the uniform flow along a flat plate. The velocity distribution is expressed by the polynominal of the distance from the wall, and the resistance of the plate of length l is obtained as [numerical formula].

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It is the well known phenomena that the efficiency of full scale water turbine is better than that of model. And it has been considered as such only because of the difference of friction losses between full scale and model, by their difference of Reynolds' number. The author, considering the discrepancy of velocity diagrams between full scale and model, owing efficiency difference, and to arrange this difference to the efficiency calculation formula, deduced a new formula to calculate the loss independent of Reynolds' number. By this formula, considering also the difference of friction loss, the author presents a new type efficincy calculation formula of Francis water turbines by their model efficiency. It is applicable to the case of partial load, and involve the deviation of maximum efficiency condition between full scale and model. The several actual experimental results in Japan coincide closely with calculated full scale efficiency by above formula.

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Aiming at the energy of wake circulation, the author deduced a new type efficiency calculation formula of propeller water turbines. Adding to this formula, assuming that in the velocity diagram, the inlet absolute velocities arrow heads always lie on one fixed straight line, and considering from the side of loss independent of Reynolds' number, the author deduced another efficiency calculation formula for propeller water turbines. But, it turned out exactly the same formula with Francis water turbines. By courtesy of The Factory Hitachi, comparing the full scale experimented efficiency of Kaplan turbine of The River Nippashi, experimented as a propeller turbine, with the calculated efficiencies by the above two formulas recognized close agreement with each other. But, up to maximum efficiency, former formula was better than the latter, and as a whole the latter was better than the former.

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In this paper, the flow pattern in the channel of a centrifugal impeller is analysed by the co-ordinate fixed to the rotating impeller. The method of the analysis is considered both in the case of the two dimensional flow and in the case where the width of the channel is varied.

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We consider a vane-wheel consisting of a number of radial vanes placed at equal angular intervals and symmetrical about the center of the wheel. The socalled virtual mass of the vane-wheel, when it makes torsional vibration about its center, and immersed in water, has been calculated by means of a two dimensional potential flow theory. Also it was shown that, when a body immersed in water it vibrates and at the same time the flow around the body takes place, the effect of the surrounding water upon the vibration of the body appears in two ways, viz. ; -(a) as virtual mass, (b) as damping force.

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On the theoretical analysis of the axial flow pump, it has been assumed that all stream lines are on cylindrical surfaces, the axis of which is the axis of rotation. But apart from the normal condition this assumption will not be correct. The fluid gets together near the boss when the discharge exceeds the normal one, while near the tip when the discharge is less than the normal one. Particularly, when the discharge is less than a certain value, the back-flow appears, and this condition gives a marked effect on the characteristics of the axial flow pump.

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The theoretical characteristics of the friction pump was already studied by the author, considering the flow in the pump as laminar. The case of turbulent flow was taken up, in the present paper, to contrast the characteristics in two modes of the flow, and to inquire into the effect of turbulence on the characteristics of the pump of improved type, already given by the author. It has been concluded that : (1) The turbulence improves, in general, the characteristics of friction pump, whatever it may be, the improved type or otherwise. (2) As to the usual type, the maximum efficiency is found to be 2/5 for turbulent, compared with 1/3 for laminar. (3) The turbulence does not change the fundamental principle for improving the friction pump, already pointed out by the author in regard to the laminar flow.

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In order to design an Westco-pump, applying the theoretical formulae derived in last reports, the estimation of the turbulent coefficient k is required. In this report, several factors which affect the value of k are discussed both theoretically and experimentally. Firstly the effects of size and form of impeller and water passage on k are determined by experiments. Secondly the theoretical formula of internal leakage is derived and verified experimentally, since this quantity also affects the apparent value of k. Thirdly on the torque of the pump, the author applies the theoretical equation to many experiments and estimates the values of correction coefficient, which is a function of the apparent value of k. Lastly the cavitation of pump is studied. which occurs at high speed and high suction head of pump.

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Previously a certain person designed the Westco-pump on his experience and flair only, but by means of this research we have now a rational way to design this type of pump. Being able to explain the characteristics of the pump according to preceding researches, the author then develops the former results in a convenient form to the practical design of the pump. He describes the values of some coefficients in this report which are contained in the theoretical formulae, and thus we can easily design the pump of maximum efficiency at the rated condition. Finally he gives an example of design.

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By considering the phenomenon of viscosity of oil, the authors obtained the theoretical equations of driving power and total efficiency of gear pump. Judging from the experimental results, it was proved that theoretical equations are exact.

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The best values of the side clearance s and the tip clearance δ of gears are given by the following equations ; [numerical formula] where C_s and C_s are constants determined by the construction of gear pump. μ=viscosity of the oil, N=no. of revolution, and P=pressure given to the oil by the pump.

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The author calculated theoretically the pressure distributions at the side clearances of gears and the quantity of leakage through there per unit time for the gear pump of a type different from that had been treated in 1st and 2nd reports, and the results were proved by experiment.

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The 1st paper dealt with a study of laminar flow of fluid in the clearance space between the gear tip and the casing, and gave the optimum gap of clearance suitable for the minimum shearing resistance to the motion of gear tip and the maximum discharge through it. The 2nd paper treated the turbulent flow, which may be set in the clearance by the intermittent sweep of a gear tip. It was concluded that : (1) If the mixing length of turbulence is unchanged irrespective of the change of clearance, we have the similar optimum clearance. (2) If the mixing length is changed in proportion to the gap of clearance, the result becomes different. We have such an optimum suitable for the maximum ratio of the water power given by the clearance and the required power for driving the gear tip.

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In the first step, author gave the methode which was obtained by solving the fundamental equations (7), (8) graphically, to get the correlation between the operating force and speed of surbo-pistion and the removal of the controller valve of three gears pump principle automatic controller. Then we compared the experimental valve obtained from the data about the stability of Neufeld Kunke valtage regulator with the theoretical valve calculated by means of above methode. Result was coincided as shown as in Fig. 11.

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There are Schiller's and Boussinesq's solutions as mathematical solutions for the velocity distribution and pressure gradient in the inlet of a circular pipe for laminar flow. But they are not satisfactory solutions for over the total range of inlet length of laminar flow. We have operated, this calculation by the numerical solution which has often been used by Asawo. When this resultis compared with the Nikuradses experiment, these velocity distributions coincide over the total range except a particular part. Moreover we have got z/aK=0.24 as a inlet length of laminar flow in circular pipe, when we admit the error 1/200 as velocities and 1/100 as pressure gradients.

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Given a thin cylindrical shell of finite length, it is initially under the action of a slight pressure (q_o) on its wall face, and making slight initial deformation (for ex., deformation due to its own weight) After the instant t=0, the cylindrical shell is subjected to a uniform external pressure q (whose absolute value is large in comparison with q_o). Then the wall of the shell begins a vibratory motion. The Author has made theoretical calculation of the value of the ratio of amplitude of vibration to amount of initial deformation of the shell. This ratio is thought to represent a scale of the effect of suddenly applying external pressure q to a cylindrical shell. In order to facilitate actual estimate, practical approximate formulae, charts, and tables of numerical values of coefficients involved are given.

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The Author has made a theoretical study on pressure-rise in a penstock, which will occur when the down stream end is closed gradually. Following two items are treated in this first report : (a) effect of tail pipe (such as a draft tube of water-turbine) upon the pressure rise. (b) effect of rate of closure not being uniform, especially slight convexity or concavity of time-closure curve, upon the pressure rise.

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When a valve, which is installed at the end of a pipe line, has a tendency to decreaseits opening with the increase of the pressure difference, a growing or sustained water-hammer will be generated in the pipe line. In this case, two series of waterhammer are possible : one has a longer period than the fundamental natural period of tae water column, while the other has a shorter period than the latter.

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Studying the grinding characteristics of the bituminous coal (Bibai-coal, 20M) by a newly designed voltex pulperizer (closed circuit type, dia. 58.6, height 177), we found that the size distribution of the ground product holds for the exponential law of Rosin and Rammler (R=100e^-bxn), concerning the particle size distribution, as that of grinding by a ball mill, and the value of exponent of n is nearly 0.44, while the value ground by a ball mill is about 1.0. Further, the value of b was nearly proportional to the applied energy and was inversely proportional to about 0.7 power of charging rate of grinding particles, so far as we experiment up to theoritical energy of air jet 70kg·m/s and discharge of air 8×10^-2kg/s. Then, the grinding characteristics was pretty obvious, from which we have been able to derive the design of the apparatus and its operations.

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