Category Archives: Sonar-Collecttors

Technical solutions adopted for the house

2.1 Materials

Since the area is seismic, the most common structural technology used in new buildings is the reinforced concrete. However, this technology is not very ecological because a great amount of energy is needed to produce cement and the difficulty to recycle the materials when its life cycle is over. Single layers of local stone and in some cases bricks constitute the structure of the old buildings.

Fig. 7 — South elevation

The house in Casaprota is planned to be built with structural walls seismic resistant made of high-energy performance bricks whose properties include high thermal inertia and insulation properties (lambda = 0,12 W/m K). The external walls made of a single layer of this type of brick glued together with special insulation plaster will ensure very good energy performance and avoid thermal bridges (U-value 0,22 W/m2 °C such as normal 25 cm thick Poroton brick with 10 cm thickness of external thermal insulation). This simple solution can be easily disseminated among all the local contractors.

As to the windows, special low-e double glass argon filled with wood frame is used. Pavement and roof have also low U-values, as shown in Table 1.

Overview of Results

As a combined effort, the refurbished building has effectively reduced electricity consumption and so not only costs but greenhouse gas emissions. Had the Council simply rebuilt without taking into account the ESD issues employed, consumption and greenhouse gas emissions would have tripled, and in the long run have also cost them more. Cowra Shire Council made a conscious effort to set the benchmark as leaders in local government, practicing what they preach for ratepayers to reduce costs and energy

use. The direct tangible benefit, of course, is the reduction in overheads for the Council. But the long term benefits to the Council and Cowra of being environmentally responsible also engenders good community incentive and positive gain to all.

The completed building has proved to be an asset to the Council from several aspects —

• the improved work environment has resulted in increased productivity and morale from staff;

• the use of solar passive and energy efficient methods has resulted in improved running costs; and

• the implementation of many of its special features has enabled Council to put into practice issues of being environmentally responsible.

To some, it is simply a newer, nicer building. Bright colours, open spaces and a pleasant work environment are all pluses. On the surface, the building is not that much different to any other building. But whilst not necessarily being outwardly different in issues of being a ‘green building’, the resultant building shows that there are many small yet easily achievable ways to be green. That it is not hard, and is in fact beneficial, to reduce energy consumption and overheads, take advantage of the sun (despite the poor orientation of this building) and take on environmental responsibility.

The building is ecologically, socially and fiscally responsible. Its many ESD components reinforce that environmentally sound buildings can still look like ordinary buildings. The Cowra Admin Building has won several awards, mostly in recognition of its ESD components, and continues to deliver to the Council a friendly, efficient and safe work environment, in which the staff and public alike are proud to have invested. It goes to show that ESD doesn’t have to be a weird word applied to weird buildings.


1. Information provided by WaterFurnace Geothermal Engineering, Wayville SA;

2. Figures based on actual kWh electricity consumption, provided by Advance Energy & Country Energy accounts for periods indicated;

3. Based on figures provided by Dr Paul Bannister & SEDA’s Building Greenhouse Rating;

Thermal analysis within a duct

The optical analysis carried out through ray tracing techniques should be necessarily complemented with a thermal analysis of the duct to assess the eventual impact in terms of wall and air temperature, and radiation losses. A numerical analysis was performed to evaluate the forced convection heat transfer characteristics of air flowing through a rectangular channel with hydraulic diameter (pitch length) of 2 mm, length 100 mm and

0. 1 mm wall thickness under laminar flow regime (Reynolds number near 6). Wall and
fluid temperature and velocity profiles were analyzed for different profiles of radiation impinging onto the channel walls. A 3D numerical model was developed with the computer code FLUENT®. Heat transfer equations were solved simultaneously in the fluid and in the solid phase. The simulation is described with the following assumptions:

— The transport processes are considered to be steady-state and three-dimensional

— The flow is incompressible and laminar.

— Thermal forced convection and thermal radiation into the channels.

— Air physical properties are temperature dependent and solid properties are not temperature


Boundary conditions for the fluid are: Constant mass flow, temperature and pressure at the inlet surface. Boundary conditions for the solid are: External walls adiabatic. The internal walls (in direct contact with the fluid) are considered as a heat source surface for different heat flux profiles as predicted by the ray tracing software. The channels were modeled with tetrahedrons (1 million for the fluid and 0.9 millions for the solid). Steady state solution of the laminar Navier-Stokes equations with conjugated heat transfer calculation was done. Inlet mass flow of 3.6.10-6 kg/s (0.77m/s) at 300 K was considered. Physical constants of the solid material were considered not temperature dependent (202.4 W/m-K thermal conductivity, 871 J/kg-k specific heat and 2719 kg/m3 density).

The air is considered as incompressible gas with physical constants as specific heat, thermal conductivity, viscosity and thermal diffusivity being temperature dependent.

Fig. 5. Inner channel heat flux (W/m2) profiles at 250, 500, 1000 and 2000 suns concentration on aperture area (1 sun= 1kW/m2).

Four different heat flux profiles on the wall were simulated corresponding to inlet aperture values of 250, 500, 1000 and 2000 kW/m2 respectively. Fig. 5 shows the heat flux distribution for each case. For the sake of simplicity we assume for all the cases equal heat flux profiles in the four channel walls (top, bottom, left and right). In addition, other cases analyzed were a wall with constant heat flux distribution from 1000 kW/m2 on channel aperture area for two different channels pitch (2×2 mm and 5×5 mm).

The consequences of the flux distributions abovementioned are represented in Fig. 6 and Fig. 7. As it can be observed in all the cases the temperature profiles are quite similar and only the absolute values differ, with the exception of the flat flux distribution.


250 kW/m2 500 kW/m2^^1000 kW/m2 — — -2000 kW/m2 Smooth 1000 kW/m2|

Fig. 6. Axial profile of air temperature at the center line for different heat fluxes.

0 0.02 0.04 0.06 0.08 0.1 0.12

Position (m)

250 kW/m2^^500 kW/m2 ■ ■ -1000 kW/m2 — -2000 kW/m2^— Smooth 1000 kW/m2

Fig. 7. Inner wall temperature through the channel for different heat fluxes.

Fig. 8. Wall and air temperature for a heat flux of 1000 kW/m2 on aperture (left) and ideal volumetric receiver behaviour (right).

The typical distributions of air and wall temperatures are represented in Fig. 8. When photon penetration takes place in the first 5 to 20 mm with the sharp profiles predicted by ray tracing techniques, the fluid gets a dramatic increase of temperature just at the entrance. The high concentration of photons at the front edge leads to some increment of wall temperature as well. It is therefore clearly observed that the limited penetration of photons inside the channel represents also a limited volumetricity of the absorber since the high temperatures are obtained too close to the inlet surface. Wall temperatures have also the maximum value extremely close to the edge. This behaviour differs from the theoretical performance of an ideal volumetric structure as depicted in Fig. 8. With those profiles radiation losses can raise up to non-negligible values of 7% (See Table 1).

The flat flux distribution is represented only for 1000 kW/m2 on the aperture, to visualize the dramatic effect on temperature profiles when a more volumetric behavior is given. For a highly porous absorber with specular surface, a more even distribution of photons would be obtained and then air and wall temperatures will show a continuous and steady increment throughout the whole channel. Maximum temperatures are in this case predicted at the channel outlet, temperatures are higher and radiation losses significantly lower (Table 1).

Inlet Flux on aperture area (kW/m2)







Radiation Losses (%)






Average fluid outlet T(K)






Average Wall T (K)






Table 1. Summary of simulation results with FLUENT

B Fluid-01 * Wall-01

Fig. 9. Air and wall temperature distribution for absorber reflectivities of 0.1 and 0.7, for specular surface, 2 mm pitch length and 30° viewangle.

Based on the previous results we may confirm that Fluent® is a useful tool to determine the relevance of design parameters like the distribution of radiation flux inside the channel to quantify efficiencies and working temperatures. Since absorber reflectivity is considered one of the key factors regarding internal distribution of photons, we have analyzed in detail two different channels with reflectivities of 0.1 and 0.7 with penetration profiles as presented for the specular model in Fig. 2. The thermal analysis has been conducted for the specular case since it shows the higher photonic penetrability.

The plots depicted in Fig. 9 explain the reason for the higher radiation losses in lower reflectivity materials. As it can be observed, higher wall fluxes are obtained at the first part of the channel with the corresponding higher wall temperatures what leads to an increment in radiation losses2.

In these cases only the radiation absorbed on the first 50 mm of channel walls has been taken into account what is enough to represent the radiation losses through the aperture.

Absorber reflectivity



Radiation Losses (%)



Average Wall T (K)



Table 2. Calculated radiation losses and average wall temperature for the first 50-mm channel and absorber reflectivities of 0.1 and 0.7as described in Fig. 9.


To assess the volumetricity and adequacy of a porous absorber as a 3D black body is not a trivial task. The optical and thermal performance of a monolithic structure are consequence of a number of design parameters like optical view angle between concentrator and receiver, absorber material reflectivity and specularity, channel pitch length, and other. The combined effect of those parameters needs to be known as it is clearly evidenced in the present analysis.

The proposed methodology integrating ray tracing techniques to determine the inner flux distribution throughout the channel and a CFD code for heat transfer simulation, reveals as a useful procedure to discern pros and cons of design parameters.

Even distributions of solar heat flux inside the channel, or in other words higher photon penetrations, lead to a higher volumetricity and more efficient design. This effect may be achieved working with larger pitch lengths and with lower view angles, but pitch cannot be indefinitely increased without penalizing radiation losses.

On the other hand, highly reflective and diffuse materials lead to counterproductive results since the penetrability improvement is neutralized by the increment in reflection losses at the channel entrance.

The results obtained under the present simulation confirm the difficulties to develop a monolith with operational behaviour close to that one from a theoretical volumetric absorber. The large gap of thermal efficiencies noticed nowadays at prototypes tested under real solar conditions and a theoretical volumetric receiver can be explained according to the previous contradictory effects predicted by the simulation.

Description of standard DEC cycle employing rotary dehumidifier

The majority of the DEC systems employing solid sorbent include a rotary dehumidifier. The plant’s components are installed in an air-handling unit and are activated according to the operation mode of the air-conditioning system. These operation modes implement different physical processes for air treatment, depending on the load and the outdoor air conditions. The most commonly used desiccant cooling process, works as follows (see Figures 1 and 2): The ambient air (1) is dehumidified by means of a desiccant wheel; the air temperature increases, since the process is nearly adiabatic (2). The heat recovery wheel is used to cool


Figure 2 — Simplified scheme of a standard DEC air handling unit

the supply air by means of indirect evaporative cooling (3-4). Depending on the air inlet temperature and humidity supplied, the temperature is further reduced by direct evaporative cooling in the humidifier, with a simultaneous increase in humidity up to condition (4). The coil on the supply stream is in operation only for heating conditions. The fan releases heat, leading to an increase in the temperature of the supply air, which brings about the supply air condition (5). An increase in temperature up to 1°C is usually expected after the considered by fan.

The return air from the building is in state (6). The air is then humidified as close as possible to saturation (7). This state guarantees the maximum potential for indirect cooling of the supply air stream through the heat recovery wheel. The heat transfer process from (7) to (8) leads to an increase in the temperature of the return air stream. The air is subsequently reheated in the heating coil until it reaches state (9). The temperature of the latter is adjusted such as to guarantee the regeneration of the sorption wheel (9 to 10).

Pre-sizing of the solar system

Numerous ‘rules of thumb’ exist to asses roughly the size of the collector array, basic cost figures such as cost of produced solar heat and the like. They are useful to obtain a first idea of the system size or to pre-select the type of solar collector, but they have to be handled with care, since they often do not consider any correlation between solar gains and load structure and neglect many important boundary conditions. Nevertheless, one of the pre­sizing methods shall be illustrated here, as it estimates the required size of the solar system (in terms of solar coverage of the cooling load) in order to obtain primary energy savings.

The specific primary energy consumption of a solar/fossil fuel driven hybrid system is defined by expressing the amount of consumed primary energy per produced kWh cold:

PE = (1 — SFcool) + pe, with

spec, sol spec, aux

efossil ‘ COP

SFcool: solar fraction (0 — 1 = 0% to 100% coverage of the cooling load by the solar system);

£fossil: primary energy conversion factor of a burner using fossil fuels;

PEspecaux: specific primary energy consumption of auxiliary components,
like for example a cooling tower and fluid pumps;

COP: thermal Coeff. of Performance = (delivered cold) / (driving heat for the cooling process)

In the following example shown in Figure 2, the specific primary energy consumption of a solar assisted air-conditioning system is compared to the corresponding primary energy consumption of a conventional ‘reference’ system, which is here assumed to be an electrically driven compression chiller system.

The thermal COP is varied from 0.6 to 1.2, covering thus the nominal COP’s of adsorption chillers (0.6-0.65) and of absorption chillers (0.7 for single-effect chillers and 1.0 — 1.2 for double-effect chillers). More information on current available air-conditioning technologies may be found in /2/.

For the conventional system, the COPconv is defined for an electrically driven compression chiller as COPconv = (delivered cold) / (electricity demand of the chilling system), and two COPconv values, namely 2.5 and of 4.5, are included into the figure.

The curves in Figure 2 are drawn as a function of the solar fraction. It can be seen that in case of a reference system with a COPconv of 2.5, a solar fraction of at least 0.38 is required if the thermally driven system operates with an average COP of 0.7, to match or fall below the primary energy consumption of the conventional system. In case the reference system works with a COPconv of 4.5, for the same solar assisted thermally driven system a solar fraction of > 0.7 is required. The curves in this example are calculated using the following conversion factors: efossii = 0.9 (kWh of heat per kWh of primary energy), and a primary energy conversion factor for electricity of 0.36 (kWh electricity per kWh primary energy).

Figure 2: Specific primary energy consumption of solar assisted cooling systems with different thermal Coefficients of Performance COP as a function of the solar fraction.

The specific primary energy consumptions for conventional electrically driven compression chillers, COPconv, are shown for two assumed values (horizontal lines).

World Wide Best practices for building integration of solar heating systems

Lex Bosselaar, Novem; Li Hua, Novem; Bart van der Ree, Ecofys; Tjerk Reijenga, Bear Architects; Alex Westlake, IT-power; Zhu Junsheng, CREIA; Yang Jinliang, Tianpu; He Zinian.

Most solar water heaters are still installed as an ad-hoc addition to a building, but the trend is to integrate the collector and the rest of the system into the building. The reasons for building integration are improved architecture, price-reduction and improved quality. In China, the world largest market for solar heating products, building integration is seen as a major bottleneck for further development of the market especially for new buildings. The United Nations Department of Economic and Social Affairs (UNDESA) with financial support from UNF and UNFIP, initiated a project to overcome this bottleneck. One of the activities in this project was a study

[1] to collect the best practices for building integration world wide. This paper is based on the work performed for this study. The conclusion is that currently most installed collectors are still not building integrated, but building integration is essential for further market growth not only in China, but also in the rest of the world.

World market for solar heating

The markets for solar water heaters are typically local markets. Most products are sold in the country where they are produced. Table 1 gives an overview of the world market, based on reference [2] and [3]. The table is limited to glazed solar collectors, including evacuated tube collectors. This table covers only about 50% of the world population, but about 90% of the world market for solar collectors. It shows that China is the leading market in the world with 10 million square meters per year, which is nearly 78% of the yearly installed collector area in the world. Looking at the installed base, it shows that about half of all the solar collectors in the world are located in China.

Table 1: Overview of the world market for solar collectors (ref [1,2,3]).

Installed million m2

Installed in 2002



Population in million













USA, Can., Mex.















n. a






n. a


1 000


1 ,33





World market





Solar thermal properties

The active thermal absorbers for water carried heat serves three purposes; delivering heat for domestic hot water and possibly also for space heating, reducing the heat load in the interior during summer and cooling the photovoltaic cells in order to increase the electrical efficiency. The full scale window prototype was used in indoor and outdoor measurements for determination of the incidence angle dependency and
the U value of the thermal collectors. Based on these results the annual energy yield has been derived.

Figure 7. The U value of the thermal collector as a function of AT.

To estimate a U value for the prototype of the solar wall operating as a solar collector, measurements of the heat loss from the collector have been performed in a dark surrounding at different temperatures of the inlet water. The values of U0 and Ui were estimated to 4.0 W/m2K and 0.046 W/m2K2 and the resulting collector U value as a function of AT is shown in figure 7.

At AT = 30 K, the U value is 5.4 W/m2K per glazed wall area. During the measurements, the prototype was surrounded on both sides with the same temperature. In a building, however, the back side of the window will usually be surrounded with air of room temperature, which will suppress the heat losses. Approximately 10 % of the total heat losses, are estimated as border losses.

The indoor measurements have been performed by using a large solar simulator providing nearly parallel light and adjustable for solar altitude angles (Hakansson H. (2003 a and b). As described in Hakansson (2003 c) and Gajbert et al. (2004), the simulator provides relatively good parallel quality of light, though this has been achieved somewhat at the expense of the light distribution over the area. As different solar altitudes are simulated, there is a tendency of varying irradiation over the test area as the uneven light pattern moves. To continuously measure the variation of the total irradiation on the test area, a number of parallel connected photodiodes were evenly spread out over the front glass of the prototype, giving a current corresponding to the received total irradiation (Gajbert et al, 2004).

By indoor measurements of the thermal efficiency, the incidence angle dependence was derived. The prototype of the solar wall was placed perpendicular to the solar simulator and the simulator was raised in order to simulate every tenth degree of solar altitude angle, i. e. the incidence angle on the glazing projected in the transversal plane, 0T. The optical efficiency, i. e. the zero-loss efficiency, of the prototype was calculated for different solar altitude angles, giving the incidence angle dependence for transversal angles, qT (QT ). The result was normalized by an outdoor measurement where the absolute value of the optical efficiency, at 30° incidence angle was registered.

The measured zero-loss efficiency has been divided by f(Qi), the transmission of the glazing, as described in equation [3], resulting in a graph angular dependence of the reflector only, R(0r). These functions, R(0r) and f(Q), shown in figure 6, were used in MINSUN to simulate the annual thermal energy yield. The measured values of optical efficiency at high incidence angles are less reliable, due to the non-uniform light distribution on the small projected area. Therefore, for angles higher than 50°, the same theoretical values as were used for the electricity calculations, shown in figure 6, have been used. The difference between the theoretical graph of optical efficiency, calculated by ray tracing, and the calculated values is due to the reflections on the absorber, the collector efficiency, the multiple reflexes on the reflector at higher incidence angles, and on the unevenness of the reflector. Since the photovoltaic cell situated in front of the thermal absorber has an efficiency of 15%, only 85% of the in MINSUN simulated irradiation falling onto the absorber is taken into count for thermal energy yield.

The result of the simulations shows that the annual thermal energy provided by the Solar Window is 103 kWh/m2 glazed surface, calculated at an operating temperature of 50°C. For operating temperatures of 40°C and 25°C the yield would be 155 and 250 kWh/a, m2 respectively. However, as previously discussed, the real heat losses would probably be lower, because of the higher indoor temperature, thus implying higher yields.

The need of automatic control systems for solar plants

One of the main challenges in power plants and particularly in those using solar energy is the increase in efficiency, trying to operate the facility in those operating points in which, the collected energy coming from the sun (which magnitude cannot be controlled), maximizes the output of the installation (net energy, in general). As solar plants are intrinsically time-varying systems, the maximization of their efficiency leads to continuous changes in the operating conditions, thus being difficult to operate them manually and requiring skilled operators. Moreover, the process efficiency will in general depend on the state of the process and the exogenous disturbances (mainly solar radiation evolution) and thus it is difficult even for an expert to predict all the time what is the best choice to maximize efficiency and profit without incurring in operation risks.

Automatic control systems help to cope with this kind of problems. The implementation in a single computer or in a computer network of suitable control algorithms aimed at controlling
the plants while optimising efficiency and profit and minimising risks is one of the main objectives of the work that is currently being performed at CIEMAT-PSA, that is briefly summarised in this paper.

Compressor and control

The first prototypes were equipped with a standard Danfoss BD35F direct current compressor and an external electronic control. A big electrical capacitor (60 mF) was used in order to overcome the start torque.

During 2003 a quite new compressor BD35K became available. The new compressor is using R600a (isobutane), which does not contribute to the greenhouse effect. A new integrated electronic control was also available. This control has been developed to ensure that photovoltaric solar panels can be connected directly to the compressor without an external control and/or capacitor. The compressor is able to do a smooth start at low speed and is equipped with an adaptive energy optimiser (AEO-control). By using this control, the compressor will slowly speed up from minimum to maximum speed (from 2000 to 3500 RPM). If the panels can not give sufficient power, the compressor will stop and after a short while it will try to start again. If the start fails, the compressor will try to start again after another one minute. Once the power from the solar panels is sufficient, the compressor will start at low speed and slowly speed up again. The controller accepts a voltage between 10 and 45 Volts. The voltage from solar panels can vary, so this new feature is good for solar powered refrigerators and freezers. On a 12 V module, the compressor needs a current of about 4,5 A to start, and it can run continuously at 2 A.



BOS components

Normal solar refrigerator


Cables, charge regulator, blocking diode



Cable (with plugs)


Photol: Prototype of vaccine cooler. The vaccine will be placed in three baskets, placed vertical in the left side of the cabinet. The ice storage is placed under the blue lid in the right side of the photo. The compressor is placed in a room under the ice

The vaccine cooler cabinet was build by Vestfrost, and is based on a highly insulated standard cabinet. The net volume of the vaccine compartment is about 50 litres and is separated from the ice storage of about 18 kg, made by a number of standard plastic containers. The evaporator is integrated into the ice storage end during daytime forced convection is cooling the vaccine. If the temperature in the vaccine compartment gets to cold during daytime, a small electrical heating element is keeping the vaccine above freezing temperature. A thermostat controls the heater. During night time the vaccine is kept cool by natural convection from the ice department.

Result of the Calculation

Fig. 8 shows the integrating value of air-conditioning load on the hottest day in Tokyo when p*=0.1~0.9 in different insulated states of the rooftop slab. Fig. 9 shows the convective heat flux from the rooftop during the daytime under the same conditions. When the rooftop is insulated, the air-conditioning load decreases (Fig. 8). The thermal resistance level of the rooftop slab was particularly conspicuous in its drop rate between no insulation and the next-generation energy-saving level of Japan (2.29m2K/W). Even in the building without insulation, by implementing a rooftop cooling of p*=0.9, the effect of air-conditioning load reduction corresponding to that in the next-generation energy-saving level could be obtained. By comparison, as the thermal resistance value of the rooftop rises, the convective heat flux from the rooftop during the daytime becomes greater (Fig. 9). However, that increase rate was merely 2~3%, a small figure in comparison to the air-conditioning load.

Fig. 10 shows the relationship between the daytime and nighttime rooftop surface temperature and the heat flux to the indoor side and the outdoor side air on the hottest day in Tokyo. In the daytime, there was a correlation between the p* value and the rooftop surface temperature; as the p* value increased 0.2, the surface temperature decreased 5.3 ~5.5°C (Fig.10(a)). The rooftop surface temperature greatly changed depending on the p* value rather than the insulated state of the rooftop slab. When p*=0.1, the rooftop surface temperature was 21.0°C higher than the outdoor temperature. As the surface temperature increased, so did the heat flux to the indoor side and the atmosphere. Yet, by strengthening insulation, the heat transfer could be controlled to a low level. Similarly, at nighttime, as the p* value increased, the rooftop surface temperature decreased (Fig.10(b)). It is believed that the increase in the p* value helped the effect of inhibiting the daytime thermal storage occur.

Thermal resistance of roof slab [m2K/W]

Fig.9 Convective heat flux to the atmosphere during the daytime (Tokyo, the hottest day, 10:00-15:00).

Surface temperature [°С]

Heat flux to the atmosphere Heat flux to indoor

Ф, (dotted line): Decreasing rate of air-conditioning load of the day due to the increase of the thermal resistance of roof slab, comparing to the condition of no thermal insulation

30 40 50 60

Fig.10 Average rooftop temperature during the daytime and the nighttime, and heat flux to indoor and the atmosphere (Tokyo, the hottest day, daytime: 10:00-15:00, nighttime: 22:00-following 3:00).

Just as in the daytime, as the surface temperature increased, so did the heat flux to the indoor side and the atmosphere. Yet, when the surface temperature went below the outdoor air temperature, the heat flux value to the atmosphere side became negative. Thus, the heat flux to the atmosphere was negative when the p* value was 0.7 or greater regardless of the insulated state of the rooftop slab. The above findings clarified that the rooftop surface temperature was closely related to both the heat load to the atmosphere and the indoor heat load. This suggests that observed values of the rooftop surface temperature in Fig. 4 had a significant role in measuring the thermal performance of the rooftop slab.

Fig. 11 shows the daily integrating value of air-condition load on the hottest day when w and p changed to 0.1~0.9 respectively. It became clear that even when the rooftop slab was not insulated, an energy-saving performance similar to that in the next-generation insulation standard of Japan (thermal resistance=2.29m2K/W) could be obtained from p: 0.5

or greater (w=0.0) and w: 0.1 or greater (p=0.1). It also became clear that, with p: 0.5, an effect of air-conditioning load reduction similar to that with w: 0.1 could be obtained. Also, regardless of the insulated state of the rooftop slab, it was confirmed that with an increase in p, w would contribute to the reduction in the indoor air-conditioning load.

Up to this point, we have examined the cooling effect of the rooftop surface in summertime. In winter, the reverse effect such as a heating load increase is believed to occur. Therefore, we made an annual load calculation targeting the regions having different climate conditions, conducting a yearlong evaluation. Fig. 12 shows the integrating value of the region-by-region annual air-conditioning load on the non-insulated rooftop slab when p*:

0. 1, 0.5, and 0.9. The greater p* was, the cooling load became smaller and the heating load increased. This problem of the heating load increase can be resolved by passive actions such as watering only in summertime using water-retentive materials on the rooftop. The details, though, will be a future task. In Tokyo, the cooling load which had been reduced due to the rise in p* was offset by the increase in the heating load and showed no change as the annual air-conditioning load. However, the problem of the peak electric power in summer was one of the major tasks in terms of energy issues. Thus, we believe that, in regions south of Tokyo, one should strive to decrease the surface temperature on the rooftop surface.


In this study, we conducted an outdoor experiment using test pieces with a focus on the passive cooling on the rooftop surface in summertime to clarify the cooling effect of various kinds. We also clarified the influence of the rooftop cooling upon the indoor heat load and heat load on the atmospheric side. The findings are as follows:

1) The results of the outdoor experiment verified that one needs to actively utilize such methods as the latent heat of evaporation and the reflection of solar radiation on the rooftop surface in order to reduce the rooftop surface temperature in summer and improve the cooling effect on the atmosphere. To obtain the cooling effect on the indoor side, it is effective to use finishing materials high in heat resistance and heat capacity while at the same time blocking off as much incoming radiation heat as possible on the rooftop surface.

2) To reduce the cooling load in summertime, it is necessary to insulate the rooftop slab and cool the rooftop surface. Also, to reduce the heat load on the atmosphere, it is effective to employ passive cooling methods by which the surface temperature on the building rooftop surface can be maintained as low as possible. After showing the correlation between the summertime rooftop surface temperature and the heat flux to the indoor side and outdoor

side, we also pointed out that the rooftop surface temperature was an important index in evaluating the thermal performance of the rooftop slab.

3) We learned that, even when the insulation work has not been done on the rooftop slab, an energy-saving performance similar to that in the next-generation insulation standards could be achieved with solar reflectance of 0.5 or greater and an evaporation rate of 0.1 or greater.

4) From the calculated findings of region-by-region annual air-conditioning load, we noted that regions where the cooling of the rooftop surface would be effective throughout the year are the regions south of Tokyo.


subscript a: atmosphere subscript s: surface

subscript t: physical quantity at thickness t Qr: radiant heat transfer [W/m2]

QS: heat transfer by short wave radiation [W/m2]

Ql: heat transfer by long wave radiation [W/m2]

QV: sensible heat transfer [W/m2]

Qe: latent heat transfer [W/m2]

Qa: conductive heat transfer [W/m2] ac: convective heat transfer coefficient [W/(m2K)] ar: radiant heat transfer coefficient [W/(m2K]] aw: moisture transfer coefficient [W/(m2h(kg’/kg))] p: reflectance [-] e: emittance [-]

a: Stefan-Boltzmann constant [W/(m2K4)]

Br: ratio of emission [-]

Js: solar radiation [W/m2]

T: absolute temperature [K]

9: temperature [°C]

K: height coefficient of clouds [=0.62]

CC: amount of clouds [-]

a, b: coefficient of one-dimensional approximate equation of saturation vapour pressure [-] w: evaporation efficiency [-] f: vapour pressure [kg’/kg] l: latent heat of evaporation of water [=2512kJ/kg]

X: thermal conductivity [W/(mK)]