Recently, many computer programs of dynamic air conditioning load have been developed and used in the calculation of cooling load. However, manual methods are used commonly for practical purposes. In this paper, we have chosen an office building as the model, which is ninestory high and a basement with a total floor area of 5,300m^2. The following three manual methods, which are widely used in Japan, were checked on and the cooling loads of the above model building were calculated by those methods. (A) Takasago Method (B) Conventional Method (Inoue's Method) (C) JSHASE Method The calculation by each of these methods is done by assuming that the outdoor condition changes periodically during a day on the basis of all-daylong air conditioning. Therefore, the heat gain through exterior walls and roof is calculated by E.T.D. (Equivalent Temperature Difference) method. However, on estimating solar heat gain through windows, the following difference appeared distinctly among the methods. In the method (A), the time-lag caused by heat storage in building structure, is not considered. On the other hand, in methods (B) and (C) consideration is given to the time-lag by Carrier's Storage Load Factor Method and Absorption Modification Factor Method, respectively. Finally, the results calculated by the above three methods were compared with the result by the JSHASE algorithm based on non-steady state analysis and the following conclusions were obtained. (1) The cooling load calculated by each of the three manual methods is effected significantly by the outdoor design condition. (2) If the same weather data are adopted, there are no noticeable differences among the cooling load obtained by the three manual methods. We proposed the "Pre-Cooling Load Factor" in the case of calculating pre-cooling load of intermittent air conditioning, and illustrated the mentioned factor for the various scales and structures of office building and at the various beginning times of air conditioning system operation.

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When hot gas flowing through an inverted U passage is cooled in the direction of flow, natural draft is produced in the passage. It has been proved from the result of theoretical investigation that the draft depends on shapes of the passage, inlet gas temperature and cooling condition. Nevertheless, in a practical inverted U passage, it is very difficult to measure the draft directly. Then, in this report, the relation shown by the energy equation between the draft and the pressure loss of the passage is investigated experimentally, and the natural draft is obtained experimentally by using the energy equation. As a result of comparison the experimental results with the analytical results, it becomes clear that the both results comparatively coincide well.

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Six each undergarment for female and male, thirteen garments for female, and twenty-five garments for male are examined in a climatic chamber under the conditions of ambient temperature 20℃, relative humidity 65% and air velocity 0.3m/sec. It is recognized that approximately linear relation exists between clothing weight and its clo. value, and the thermal resistance of garment is different between the female's one and male's one. The clo. value of garment corresponding to each part of the human body has shown. It is also recognized that even if the average clo. value is similar, the partial clo. value is different, furthermore, some examples of garment expected to be comfortable under the conditions of ambient temperature 20℃ and 25℃ have shown. The heat loss from the human body wearing garments is estimated at 30〜50kcal/m^2hr.

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The heat transmitting tubes used in multitube heat exchangers are desired to have such characteristics as superior heat transmission capacity and minor draft resistance loss. However, heat transmission capacity and draft resistance loss are generally in a contradicting relation, so that the overall performance including the effect of these characteristics should be taken into account in designing heat exchangers. Hitherto, circular tubes and flat tubes have been used as heat transmitting tubes, and many fins have been attached on the gas flow side of tubes to enlarge their heat transmission area. While many reports have been published on the studies of such finned heat transmitting tubes, there are few reports systematically studied on bare or finned tubes with various cross-sectional shapes. In this paper, the effects of cross-sectional shapes of the tubes on the heat transmission capacity and draft resistance loss were experimentally investigated. Further, heat transfer coefficients of the tubes were estimated under the constant resistance horse power to drive the outer air flow, and their overall performance including two characteristics above mentioned was also estimated. It will be preferable to discuss the characteristics of the elliptical tubes in relation to those of a circular tube which is considered as the fundamental shape of heat transmitting tubes. In this experiment, six kinds of the elliptical tubes were prepared, in which a circular tube as an elliptical tube with the maximum axial ratio (ratio of minor and major axes) and an elliptical tube with the minimum axial ratio similar to that of a flat tube were included. Each of the heat transmitting tubes is of copper and has the same outside perimeter of 79.8mm, thickness of 0.8mm, and effective length of 400mm. The heat transmitting tubes were arranged in a single row with the tube pitch of 80mm. Heat was transferred in the cross flow from air of a constant temperature of 80℃ to the tube, with the air velocity ranging from 3.5 to 12m/s, and the flow rate of the water in the tubes was ranged from 3 to 20kg/min. All temperatures were measured with thermocouples and all pressures were measured by pitot tubes. The main results obtained were as follows. When the resistance horse power was low (the air velocity was low), the elliptical tube with the cross section similar to that of a flat tube was superior in the overall performance. However, when the resistance horse power was large (the air velocity was large), the excellent overall performance was obtained with the circular tube. Some problems may occur from a point of the structural strength for elliptical tube, when the pressure difference between the fluids flowing outside and inside of the tube is large. These problems, however, may be relieved by attaching many fins on the outside surfaces of the heat transmitting tube, and therefore, in case of the low air velocity the use of the elliptical tube with cross section similar to flat tube will reduce the drafting power for given heat transmission capacity, and the economical operation may be realized.

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In the first report, it was shown that the factors which influenced on the dispersion of warm water into a rectangular open channel were Reynolds number, Froude number, densimetric Froude number, the velocity ratio, the heat loss from the water surface, and several geometrical boundaries. They were derived by rewriting the equations of continuity, momentum, and energy, and boundary conditions to dimensionless forms. In the second report, a study was made experimentally to know the influence of three dimensionless factors, which were densimetric Froude number, the velocity ratio, and Froude number, on the temperature distribution, the decrease of maximum temperature, and V^* (the volume within iso-concentration contour) etc. And further the approximate equations for A^<**> (the cross sectional area within iso-concentration contour) and V^* were discussed by modifying the equations derived from the analysis of a simple dispersion model of warm water. The purpose of this paper is to investigate experimentally the influence of geometrical boundary conditions, which are the depth of conduit of warm water, the aspect ratio of conduit, and the width of open channel, on the dispersion of warm water. Ten experiments were performed in the range of D/H=0.1 to 0.8 to study the influence of the depth of conduit using a conduit of 50mm×20mm, where D is the depth of center of conduit from the water surface and H is the depth of channel water. Nine experiments and three experiments were carried out as to the shape of conduit and the width of channel respectively. The former was performed in the range of the aspect ratio B/C of 0.156 to 10 under B・C=1,000mm^2=const. The latter was performed in three rectangular open channels of W/H=2, 1.25, and 0.75, where W is the width of open channel. The measurements of temperatures were made using twenty Cr-Ar thermocouples of 0.1mm in diameter, a commercially available multipoint digital voltmeter, a micro-computer, and a perforator. The data were managed with a digital computer of high speed to obtain the co-ordinates of equi-temperature points, A^<**>, and V^*. The main experimental results are as follows. (1) The maximum temperature decreases rapidly with increase of the depth of conduit, and A_<max>^* (the maximum value of the cross sectional areas within iso-concentration contours) and V^* decrease in proportion to the depth of counduit. (2) The maximum temperature and V^* change not monotonously relative to the aspect ratio of conduit, but have maxima. A_<max>^* changes not so much when the aspect ratio is small, but decreases in proportion to the aspect ratio when it exceeds some value. (3) The width of channel influences considerably on the temperature distribution and A_<max>^* but influences little on the maximum temperature and V^* in the above-mentioned range of W/H. In addition it was confirmed that the values of A^<**> and V^* in these experiments were well expressed by the approximate equations shown in the previous report.

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Investigation of the fixture usage and water consumption in the buildings and examination for its causes will be more and more necessary for the collection of the minute designing data. By such point of view, this report is summarized for the results of investigation of the theaters in which the fixture usage is the waiting use system. The results are for the movement of the persons, frequency, duration, and simultaneous use of fixtures, partial and total water consumption in the two theaters. The auther considers that this report is the first step for the collection and examination of data on the theaters.

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In the time series response factors method, the weighting factors relating room air temperature deviation to space heat extraction are used to calculate heat extraction from cooling load. In dynamic cooling load computation method by the use of computers, heat extraction, preheating load, heat storage load and room air temperature deviation from set point can be computed by the weighting factors method in which eight terms of weighting factors and common ratio are used (WZ-WX (standard) Method). It is difficult in practice, however, to calculate cooling load by manual calculation using the same method because of a large amount of calculation process. In this report, therefore, the approximate calculation method of cooling load, in which only the first three terms of weighting factors and common ratio are used (WZ-WX (three terms approximation) Method), and the simple method, in which weighting factors and common ratio can be derived easily from graphs (Simple Calculation Methods I, II) are proposed for manual calculation. The graphs prepared for this purpose show the relationship between the ratio of total surface area of the room enclosure to the floor area and the weighting factors per unit floor area. These graphs are made using the weighting factors computed by WZ-WX (standard) Method for various types of room with different size of room space, different positions of outside walls, different values of weight of structure. In this report comparisons are made with regard to the values of weighting factors among WZ-WX (three terms approximation) Method, Simple Calculation Methods I, II and WZ-WX (standard) Method, investigation being made on room air temperature transfer function coefficients proposed by ASHRAE and theoretical consideration on the weighting factors method. The results obtained may be summarized as follows: (1) The values of heat extraction and room air temperature deviation at the starting time of air-conditioning when the values seem sufficiently converged, calculated by WZ-WX (three terms approximation) Method is compared with the values calculated by WZ-WX (standard) Method with the result that little difference is found between them. The Simple Calculation Methods, I, II are compared with WZ-WX (standard) Method by the same scheme as above and in consequence the difference between them is found by less than 5 per cent. (2) If the initial value of heat storage load is assumed to be 80 per cent of the cooling load that would occur at the same time under intermittent operation of ten hours a day, it is found that heat extraction and room air temperature deviation at the start time of air-conditioning after two days calculated by the three terms of weighting factors method are 5 per cent larger or smaller than the converged value at the same time calculated by WZ-WX (standard) Method. (3) Values of room air temperature transfer fanction coefficients proposed by ASHRAE are converted into values of weighting factors and compared with the weighting factors given by WZ-WX Method. As a result, it is found that the absolute values of the first and second terms of weighting factors are smaller and the absolute value of the third term is larger. This must be due to the difference in assumption for heat capacity of the room. (4) In WZ-WX Method, the effects of radiation interchange among the inside room surfaces are taken into consideration by the factors WX. The values of heat extraction calculated by the use of weighting factors, in which the effects of radiation interchange among the inside surfaces are considered in direct calculation turned out to be in good agreement with those by WZ-WX Method.

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Although numerous studies have been made in the numerical method for predicting room air distribution, there appear very few reports which aim to investigate its computational accuracy and performance of predicting ability by comparing the calculated predictions with the experimental ones. In our previous report, trying to supplement those deficiency of research, it was made to compare the experimental results with numerical ones especially as for room air temperature distribution and discussed the predicting ability of our numerical method including wall functions and two-equation models of turbulence which used turbulence energy and its viscous dissipation rate as two dependent variables. Following this investigation, it was tried again to compare the calculation by almost the same method as in the previous report with measurements of room air velocity and gas concentration distribution which was conducted in detail in the two-dimensional room models. Here, in this report, we summarized the methods of computation and experiment and described the results of comparison in one typical case of model configurations which had various positions of supply-exhaust opening in our experiments. As was mentioned above, the method of calculation was the same as previous one except that in this report, the concentration was used in place of temperature as a dependent variable of conserved property, but it was necessary to develop newly how to treat the boundary conditions of pollutant gas injection, then we constructed the appropriate numerical models of these situations after many preliminary numerical experiments. Experiments were made in the model room of dimensions of 32cm×32cm cross section, 64cm breadth in which the air velocity and concentration distribution were sufficiently secured of two-dimensionality by giving our careful consideration to regulate the air flow upstream of the model room. Ethylene gas was used for pollutant gas because its molcular weight was nearly the same as air's, then we could neglect the effect of density difference on the room air flow and distribution. Hot wire anemometer (KANOMAX, Model-21) and hydrocarbon analyzer (Toshiba-Beckman, Model 400-J) were used for measurements of velocity and concentration, and time averaged value of these were taken by recording the output and averaging on the charts. Following aspects resulted in this investigation: 1) As for velocity, experimental flow pattern by visualization and velocity distribution at various sections in model room agreed well with the theoretical results except that experimental curvature of jet flowing from supply to exhaust opening and maximum axial velocity of jet were showed a little difference with computations. 2) Theoretical concentration distributions gave somewhat low values compared to experiment, and difference was rather great especially in regions of high concentration distribution. This difference could be attributed to the fact that model room had a little poorer two-dimensionality of concentration in region of high value than other, and indicated the necessity of improved treatment of gas injection.

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In this report, following the previous one, we described the results of experiments of room air velocity and gas concentration distribution in the two-dimensional room models in which locations of supply and exhaust opening and gas injection opening were in different ways varied, and discussed still more the performance of our numerical prediction method by comparing these with our numerically calculated results. The details of numerical and experimental methods did not described because they were much the same as those in the previous report. The items on which we compared the calculation with the experimental results in this report were as follows; 1) decrease of velocity component on axis of supply opening, 2) variation of velocity distribution of supply air jet, 3) velocity distribution at any cross section in the room, 4) concentration distribution, 5) relationship of experimental and calculated values of concentration, 6) cumulative frequency distribution of concentration, 7) flow pattern as a whole in the room-comparison of calculated stream line with photographs of visualized flow pattern, 8) standard deviation of concentration. It was noted that although the theoretical results did not always agree precisely with experimental values, these results were satisfactory considering the present stage of development in numerical method of room air distribution.

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This paper describes a study on the evaluation of the reduction in energy consumption on various air conditioning systems. Four air conditioning systems are chosen: single duct constant air volume system, single duct variable air volume system, terminal reheat system and dual duct system. The annual heat consumption is determined quantitatively by the computer simulation for each system. It was found that the deficient heat and excess heat amounted to 20〜85% of the total heat extraction rate of a room depending on the air conditioning system and could be regarded as the indexes, which indicates follow-up ability to maintain the room temperature. The factors affecting the annual heat consumption could be identified as follows; (1) zoning, (2) the method used in determining the flow rate and the temperature control, (3) the method of inducing the outside air, (4) the setting of the room thermostat. The annual heat consumption at the air handling unit were calculated taking these factors into consideration for 26 cases, where coil arrangements, control of coil performance, setting of supply air temperature, outside air intake, etc. are varied. (1) Large differences are shown between the systemes with zoning and with out zoning. For example, terminal reheat system with out zoning consumed 2.3 times more heat for cooling, 2.6 times more for heating, than constant air volume system with zoning. These excess energy in required due to the reheat losses and mixing losses which varies according to the zoning. (2) Heat consumptions which consumed at the reheat coil to maintain the air humidity in the room are in high volume. For example, reheat system consumes 85% more heat volume than the single coil system, to maintain the room air humidity. (3) Heat consumption varies very much according to the location of the room thermostat. In the case of this study, heat consumption increased 11% under the influence of unsuitable location of the room thermostat. (4) The heat consumption of the system, which induced the outside air in proportion to the drybulb temperature, decreased by 8.5% in cooling and increased by 24% in heating. This increase is produced by humidifying load. But the heat consumption of the system, which induced the outside air in proportion to the enthalopy, decreased by 6.1% in cooling and 1.2% in heating. The heat consumption of the system with total heat exchanger decreased by 7.7% in cooling and 39.8% in heating.

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