The thermal conductivity of the air is very small. Nevertheless, when a convection occurs in the air layer, the heat is transferred by this convection, consequently the coefficient of heat transfer through the air layer becomes greater than thermal conductivity of the still air. In this report, the coefficient of heat transfer through a horizontal air layer which is enclosed and heated from below is experimentally measured, ard the experimental results are arranged by λ_LK/λ_L and G_r, where λ_LK=equivalent thermal conductivity of air accompanied with convection λ_L=thermal conductivity of air G_r=Grashoff number.

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Heat transfer resistance of the air layer is severely influenced by the state of motion in the air layer. In this report, the state of motion is observed by using both TiCl₄ and NH₃OH, and the coefficient of heat transfer through the air layer is measured as the first report. According to our experiments, the convective motion in the air layer is cellular motion and the breadth of a cell is about twice of the thickness of the air layer and the ratio between the thickness of the air layer and the breadth of a cell increases with Grashoff number. And then the heat transfer is theoretically analyzed and it is found that theoretical results agree well with the experimental data.

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In this paper, we analysed heat transmission of a flat wall, by two dimensional heat conduction theory, in the case where heat exchange occurs by heat transfer only on one side of its surface. In the theoretical analyses of heat conduction, it is usually assumed that surface heat transfer coefficient is constant. However, in actual cases, it is not constant but a function of position or temperature. In this paper, we obtained general solutions by assuming that it is a function of position, and then, showed an example of calculation by this method in the case where heat transfers to flowing air by laminar surface heat transfer. In this way, it is made clear that surface temperature distribution is remarkably different from the distribution given by an analysis of one dimensional heat conduction, if a wall in thick its thermal conductrvity is better. But there is no remarkable difference between the results obtained by both methods with regard to heat quantity exchanged on the surface.

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The effect of initial turbulence on the surface heat transfer was investigated in the case of a smooth flat plate. The initial turbulence was introduced by a screen, and the percentage turbulence √(<u^^->²)/U was measured accurately by a hot-wire anemometer in the boundary layer and in the main flow. On the other hand, the change of the temperature difference between air and a plate, during the plate being cooled, was measured in the same condition, and from its result the local heat transfer coefficient on the surface was calculated. Thus, we were able to make clear the relation between initial turbulence and heat transfer. Surface heat transfer becomes better with the enlargement of the initial turbulence and the increasing of heat transfer coefficient was remarkable within the range of small initial turbulence. In the rage of larger initial turbulence than 7∼8%, the local heat transfer coefficient increased no more, and its value was 55% larger compared with the case of the smallest initial turbulence.

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We intend to report in the present paper as to how the coefficients of heat transfer α obtained from our previous quenching experiments are applied to hardening of carbon steels. At quenching period the test bodies were rapidly cooled through the three stages, film state boiling, bubble state boiling and natural convection. Comparing the values of α_min (at film state boiling) and α_max (at bubble state boiling) with the hardenability of carbon steels, we have found that strong quenching media like NaCl aqueous solution and cold water, have higher values on α_min and α_max, and soft quenching media like mineral oil, have comparatively higher values on α_min in spite of their lower values on α_max.

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This report is one of the experimental results of heat transfer of the revolving steam condenser tube, unde the following conditions : the contained air 0∼30% in weight, and drop condensation.

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In this paper, an approximate calculation method has been tried on the characteristic curve of the steam ejector, the relation of vacuum of mixing chamber and weight of inhaled gas. And by this calculation method, the following effects have been made dear. a) The effects of each dimension of the steam ejector on the characteristic curve. b) The effects of the charaetesistic curve on the exit pressure of diffuser and the so-called hysteresis phenomena. c) The effects of live steam condition on the characteristic curve.

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In the ideal middle stages of a multi-stage axial-flow turbine, we take, as one stage, any adjoining two rows of blades such as I, II which are moving to the opposite directions with each other. The adiabatic heat drops of the moving blades I and II are (1-ρ) H₀ and ρH₀ respectively, when we represent the total heat drop of the stage with H₀ and adopt a factor ρ called "ratio of heat drops". For the blades I and II, we take c', c as circumferential velocities, ζ', ζ as coefficients of loss and β₂', β₂ as exit angles, respectively. We represent the relative velocity of fluid at the exit of the blades II with u₂. Taking the velocity ratio ξ₂= (c'+c)/u₂ and the function about energy ratio Au₂²/(2gH₀)=Φ, the circumferential efficiency η_c of an ideal middle stage may be given by the next equation [numerical formula] if we have the suitable value of the angle β₂', calculated from the following equation [numerical formula] Using these equations, we have made clear numerically the performance of double-rotation axialflow turbines. And comparing these turbines with double-rotation radial-flow turbines, we have confirmed again our old deduction. In general we may expect to realize more efficient doublerotation turbines, if we are able to get the combinations of two blades having the sections suited to the given theoretical condition, not being restrained for any stage by such construction with two rows of blades having same sections as in Ljungstrom turbine.

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The quantity of air drawn in per stroke by 4-cycle engines will be given as [numerical formula] Where, ρ=air fuel ratio, p₁=pressure at the moment the inlet valve has just closed, p_r=residual exhaust gas pressure in the clearance space, V_s=effective stroke volume, T_s=air temperature, ΔT₁=temperature drop due to the evaporation of the fuel, ΔT₂=temperature rise of the mixture, gaining heat through the cylinder walls, and other hot surfaces, R₁=gas constant, and ε^*=effective compression ratio. Then, if we obtain the values of p₁, p_r and G_a by experiment, and knowing the values of ρ, V_s, ε^*, T_s, and ΔT₁, we will be able to calculate the value of ΔT₂ from the above equation. In this way I obtained ΔT₂≒100°C by water cooled L-head petrol engine. This value shows that the heat exchange between the mixture and hot walls has a great influence upon the charging efficiency. For instance, if ΔT₂=0 in this case at n=2 000 r.p.m. the charging efficiency will increase to 0.98 from 0.72 and at n=3 000 r.p.m. to 0.75 from 0.55.

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As a compound internal combustion engine, it can be considered several combination of high pressure and low pressure cylinders. The compound internal combustion engine considered in this investigation, consists of two HP cylinders which act as four cycle engine and LP cylinder which acts as two cycle. The crank angle between HP and LP cylinder is 180° and the bottom side of the LP piston is used as a supercharger for HP cylinder. The theoretical formulas of work done and thermal efficiency are reduced and form the results of numerical calculation, the effects of ratio of stroke volume of HP and LP cylinder, volume of transfer passage between HP and LP cylinders etc. on thermal efficiency are discussed.

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We deduced the efflux gas angle from turbine nozzles and blades through simple analysis, and at the same time we knew the effect of outlet Mach number and angle on those.

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In the 2nd Report, the effect of the dimensions of swirl chamber on the size of drop was investigated and an attempt was made to search for the condition of the optimum form of nozzle. Besides the size of drop, the characters of the swirl nozzle can be represented by the discharge coefficient and the spray angle. These two can be obtained from the potential flow theory but they agree with the experiments only in the narrow range. In the present paper, the auther made the empirical formulas of the discharge coefficient and the spray angle which were applicable to wide range of the dimensions of swirl chamber for the non-viscous liquids, and obtained the more accurate empirical formula than that of the second report for the mean diameter of drop taking into accout the effect of ratio between the length and diameter of orifice.

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The all characteristics of the swirl nozzle are affected by the viscosity of liquid, but the effect upon the discharge coefficient among these is considerably complicated. In this paper, the effects of the surface tension and the viscosity on the discharge coefficient were investigated. The discharge coefficients are fairly arranged by introducing a Reynolds' number R_e of the swirl chamber, and in the range of R_e=1 600-2 800, all swirl injection nozzles have a transition from the laminar flow to the turbulent. The changes of discharge coefficients by the transition are rather gradual, including the extreme case of no change. A general empirical formula of discharge coefficient was made from the results.

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In this paper, the effects of the surface tension and the viscosity of liquid on the spray angle and the mean diameter of dop, were investigeted. The spray angle can be arranged in the main features by introducing the Reynolds' number of the swirl chamber. According to the results, we obsained a general empirical formula for the spray angle. The mean diameter of drop is directly proportional to the fourth root of the surface tension for the non-viscous liquid, The general empirical formula for the mean diamter of drop was induced from the results of the measurements, addition to that in the "Supplement of the 2nd Report". We can neglect the effect of the viscosity on the size of drop, when the dimensionless number (ρν²/σt₀)<0.015 as its effect is smaller than 1%.

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