Category Archives: Sonar-Collecttors

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.

System

Storage

BOS components

Normal solar refrigerator

Battery

Cables, charge regulator, blocking diode

SolarChill

Icepacks

Cable (with plugs)

Cabinets

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.

Conclusion

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.

Symbols

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)]

Space heating

To analyse the performance of facades addressed to produce space heating, the top part of the tank in the facade is considered linked to a heat exchanger (a thermal radiator). Water from tank is pumped through the heat exchanger delivering heat to the annex room, and then it returns to the facade entering at bottom.

The following assumptions have been considered:

-Indoor room temperature is considered constant and equal to 20°C.

-Global heat transfer coefficient of heat exchanger has been considered constant UA = 100 W/°C. Return temperature to the facade is calculated according to:

QLOAD = UAIT — Tg) = mcp(OTL — Tinlet) (1)

where:

T = 0.5 (OTL + Tinlet) (2)

Tg is the indoor ambient temperature, OTL is the outlet water temperature from the facade, is the water flow rate [kg/s], QLOAD stands for the energy delivered from the heat ex­changer and Tinlet is the return temperature.

-The facade delivers heat from 17 to 24 hours each day of the heating months.

Existing discomfort glare indices

In the past, a number of different glare indices have been developed. All of them were basically aimed for artificial lighting and considered only small sized glare sources. Only one of them, the Daylight Glare Index (DGI) has been adapted to large glare sources and daylight conditions. Velds (2000) and Gall et al. (2000) found in their tests little or no correlation between the glare formula and the user assessments. This has been proven

by our own studies in 2003, when 27 subjects in test offices at the Fraunhofer ISE were tested under different lighting conditions. As can be seen in the following graph, there exists almost no correlation between DGI and user reaction.

Figure 1: Correlation between user assessments at Fraunhofer ISE test calls, their referring physical measurements and theory for the Daylight Glare Index (DGI).

The main reason for the big discrepancy is that, for the initial study, less than 10 people were used to develop the DGI formula. Another problem was that the studies were originally carried out under artificial and not real daylight conditions, nor under real office conditions.

Chauvel et al. (1982) argued that the weak correlation between the DGI and the observed glare from windows is compounded by other visual and aesthetic factors such as the quality of the view, the appearance of the window as well as the visual and aesthetic interior qualities of the room.

There was a greater tolerance of mild degrees of glare from the sky seen through the window than from a comparable artificial lighting situation with the same value of glare index, but that this tolerance did not extend to severe degrees of glare (Boubekri and Boyer, 1992; Chauvel et al., 1982). Chauvel et al. (1982) also observed that the discomfort glare resulting from the direct view through windows has been found to vary greatly from observer to observer and also to vary with factors associated with the appearance of the window, the view outside and the surroundings.

Iwata et al. (1990/91) showed that the subjects judged the light to be less uncomfortable even after only 30 seconds, suggesting that the most serious glare problems occur during the transition i. e. the time immediately after exposure to the glare source. Also, Osterhaus (1996) observed that the research subjects (32) in his experiment commented on becoming more sensitive to glare as the experiment progressed (2-2.5 hours) and that this impression was confirmed by experimental data. Osterhaus and Bailey (1992) also pointed out that no data is currently available on perceived comfort or discomfort and the relations between comfort and task performance under conditions in which the glare source borders or surrounds a work task. All existing discomfort glare indices were developed by assessments of subjects directly viewing the glare source rather than focusing on a work task.

Osterhaus (1996) also suggested carrying out glare experiments with subjects exposed to the daylighting situation for at least the eight hours of a regular workday. Decreasing work performance would be expected due to fatigue and distraction induced by glare discomfort. Sivak and Flannagan (1991) found that task difficulty affected discomfort glare. In their study, smaller gapsizes in a gap-detection task resulted in more discomfort glare responses concerning a simultaneous presented light source. They concluded that the assessment of discomfort glare requires the inclusion of the relevant visual task the observer is involved in during the presentation of the glare stimulus.

Test Results

Data plots:

A typical testing day is characterized for a certain number of heliostats focused over the receiver. This number is increased in case of mist that makes the solar direct radiation lower, but otherwise it is fixed. On the other hand, in one test day, the mass flow rate is decreased several times until the temperatures approach certain limits in each section. In a clear day, with an approximately constant solar radiation, the temperatures increase when the air flow rate decreases. From all the data recorded for the DAS, a plot is drawn for the most relevant parameters (see fig.2): direct solar radiation (in W/m2), overall mass flow rate (in kg/s, over the secondary axis in the chart) and air outlet temperatures (in °С) from the cups of interest (all of them adjacent and located at the central part of the receiver). Moreover, the average temperature of the 36 absorber cups is represented to have an idea of the temperature level in the whole receiver.

Selection of steady states:

To present the response of the volumetric solar absorbers under certain known parameters, it would advisable that the system was in a steady state. However, this is nearly impossible, since solar radiation is always slowly increasing until noon and decreasing from then. In addition, the clouds in the sky may cause transitories in the incoming radiation. The solution is to find quasy-steady states where the main input parameters of the receiver do not change very much. But also the response of the receiver has to be considered. Thus, the input and output parameters values must keep inside a certain range during a large period of time related to the response time. This time is calculated as the time that takes the receiver to react and to reach the 65% of the final value of the output variable after a quick change in one input variable, while the other input parameters remain constant. For this receiver, the response time is in the order of some minutes, between 3 and 5 depending on the temperature selected as output variable.

GMT (hours)

To identify the quasy-steady states, several criteria have been used, but in all of them the incoming radiation and mass flow rate, as input parameters, and receiver average temperature, as output parameters, must not change more than 4% or 5% in a previous period corresponding to 4 or 5 times the response time of the receiver. In other words, to consider that a measurement value is inside a steady period it is compared to the values taken in the last minutes (between 12 and 20, depending on the selected criterion). In addition, the change gradients of these parameter must be less than 1% (or 1.25%) in every former response time step.

As an example, all the experimental points for the 3 parameters of interest that comply with the all the requirements at the same time for the one of the strictest criteria are marked in the fig.2. This criterion compared every value of direct solar radiation and overall mass flow rate with their values during the last twenty minutes and they must not vary more than a 4% (and no more than 1% every 5 minute step). Meanwhile, the average absorbers temperature must not change more than a 5% (no more than 1.25 in every step).

After calculating the quasy-steady periods of a test, the mean value and the standard deviation of each variable are calculated from its values from the beginning to the end of the steady period. Thus, the number of experimental points is reduced from several thousands to some dozens in the best days, or just a few or even any steady data when direct radiation was unstable.

MODELING OF HEAT AND MASS TRANSFER IN PARALLEL PLATE LIQUID-DESICCANT DEHUMIDIFIERS

L. C.S. Mesquita, D. Thomey, S. J. Harrison.

Department of Mechanical and Materials Engineering, McLaughlin Hall, Queen’s University, Kingston, ON. Canada, K7L 3N6. Email :mesquita@me. queensu. ca

INTRODUCTION

In the last few years there has been renewed interest in solar driven air-conditioning

[1] . Some of the work have been focused in desiccant cooling systems. Such systems have the advantage of improved humidity control, particularly in applications with high ventilation rates [2]. Most of the systems already developed employ solid desiccants, with relatively high regeneration temperatures. One alternative is the use of liquid-desiccant systems. In these systems, lower regeneration temperatures can be employed, allowing for a more efficient use of heat from low temperature sources, e. g., flat plate solar collectors [3]. Another advantage of liquid-desiccant systems is the potential of using the desiccant solution for energy storage.

The main components in a liquid-desiccant air-conditioning system are the dehumidifier and the regenerator. Many different technologies have been developed for these two components. For the dehumidifier, the most common technology employed today is the packed bed. However, packed beds must work with high

dehumidifier channel.

desiccant flow rates, in order to achieve good dehumidification levels without internal cooling. Higher desiccant flow rates imply on small changes in the concentration of the desiccant solution during the process. This, and the higher level of heat dumping from the regenerated solution that follows higher flow rates, reduce the coefficient of performance of the liquid-desiccant cycle. One option that allows lower flow rates is the use of internally cooled dehumidifiers [4,5]. Figure 1 presents the schematics for one channel of a internally cooled dehumidifier, wich is composed of several of these channels stacked together.

In the present work, mathematical and numerical models were developed for internally cooled dehumidifiers, using three different approaches. The first approach uses heat and mass transfer correlations. The second one numerically solves the differential equations for energy and species for a constant thickness film, using the finite-difference method. The third approach introduces a variable film thickness. All approaches assume fully developed laminar flow for the liquid and air streams.

Point p t h % F [kPa] [°C] [kJ/kg] [kg H2O/kg sol.] [kg sol/kg ref.] 1 7,375 40,0 167,5 1,000 1,000 2 1,497 13,0 167,5 1,000 1,000 3 1,497 13,0 2525,4 1,000 1,000 4 1,497 36,0 85,5 0,461 23,846 5 7,375 36,0 85,5 0,461 23,846 6 7,375 68,7 153,2 0,461 23,846 7 7,375 77,0 176,2 0,438 22,846 8 7,375 42,0 105,6 0,438 22,846 9 1,497 42,0 105,6 0,438 22,846 10 7,375 77,0 2644,2 1,000 1,000 Table 1: Operating conditions of the absorption chiller (for state points, see Figure 2) . Prototype building and field test

First a test prototype of the chiller with a cooling capacity of about 10 kW was built and tested under laboratory conditions. For the second prototype some improvements were done and the cooling capacity was raised to 15 kW.

Figure 3: Cooling capacity for different cold water temperatures

The cooling capacity and the coefficient of performance (COP) of the second prototype for different cold water temperatures are shown in Figure 3 and Figure 4.

The figures show a distinct dependency on the cold water temperature. The higher the cold water temperature the higher is the cooling capacity and the coefficient of perform­ance of the chiller. In this capacity range the chiller will mostly be used for cooling only but not for air-dehumidification. By considering the cold water temperature when designing the room cooling system a high COP and cooling capacity can be achieved.

Figure 4: Coefficient of performance for different cold water temperatures

Field test

After prototype testing under laboratory conditions a field test was carried out in the sum­mer of 2003. At three test sites the new absorption chiller was installed. The locations of the test sites and the different peripheral equipment that was used is specified in Table 2.

The field test showed good results. The absorption chiller worked with a high reliability and operational safety. It is able to work over a wide range of external conditions. The test in Italy showed that the chiller even works with flat plate collectors (lower heating tempera­tures achievable than with vacuum tube collectors) and a dry cooler for re-cooling (rela­tively high cooling water temperatures during daytime).

At the test site in Kothen the room cooling system (gravity cooling units without ventilation) was already installed. It is designed for lower cold water temperatures and could not be changed for this field test. Therefore the absorption chiller had to work with cold water temperatures of 10…12 °C. Also the hot and cold water flow rates were below design con­ditions. Because of these conditions the absorption chiller reached a lower COP as shown before.

Results of the chiller operation in Kothen for one summer day are shown in Figure 5 and Figure 6. On this day some variations of the solution flow rate and the desorber heating temperature were carried out.

The tests also showed that a precise adjustment of the two solution flow rates is very im­portant for achieving a high COP. If one solution pump is pumping more solution than the other one solution reservoir will frequently be empty. This results in a short stop (some seconds) of the operation of the pump. During this stop the solution heat exchanger is without effect which affects the whole cycle of the absorption chiller. The chiller needs minutes to recover and to reach the former values of operation (COP and cooling perform­ance).

Location

latitude

Heat source

Recooling

Cold water use

Neumarkt,

Italy

46,4° (N)

flat plate solar thermal collectors, 55 m2

dry cooler (fan coil)

room cooling with fan coils

Westenfeld,

Germany

50,4° (N)

waste heat of an engine driven cogeneration unit

dry cooler (fan coil)

room cooling with fan coils

Kothen,

Germany

51,7° (N)

CPC-vacuum tube collec­tors^ m2

wet open cool­ing tower

cooling of office space; gravity cool­ing system

Table 2: Test sites — location and equipment

The experiences of the field test lead to some further improvements of the chiller design to increase the COP and the flexibility. The electrical power consumption of the peripheral equipment (pumps) could be reduced.

The chiller that is shown in Figure 7 was presented at the IKK fair in Hannover in October 2003. Additional field testing and the composition of "standard” solar thermal cooling con­figurations using the small capacity absorption chiller are planned for the next cooling sea­son.

-□— temp. hot water in — — A — temp. cold water out Time

— О — temp. hot water out — O— temp. cooling water in

— V temp. cold water in temp. cooling water out

——- condensation pressure

• • • • evaporation pressure

Figure 5: Results of the field test in Kothen — temperatures and pressures (8.8.2003)

7500 7000 6500 6000 5500 4| 5000 4500 4000 3500 3000 2500 2000 1500 1000 500

ra

CL

<D

CL

Another focal point will be the coupling of the absorption chiller with other heat sources for example waste heat of thermal biomass usage or cogeneration units.

SHAPE * MERGEFORMAT

—о— Heating capacity [kW] jjme cop

— О — Cooling capacity [kW]

COP

Figure 7: Small capacity H2O/LiBr absorption chiller Wegracal SE 15

Figure 6: Results of the field test in Kothen — capacities and coefficient of performance (8.8.2003)

Подпись: COP

East-facing gallery (entrance lobby)

The main entry lobby to the museum is used as a gallery for the display of oil paintings and metal sculptures. In 1996, Philip Johnson made changes to the museum’s east fagade by replacing the glass to a dark tinted one with visible transmittance (Tvis) of less than 5% and UV protection. Even with those modifications and the arched portico, direct sunlight strikes the lobby every morning throughout the year except few days around summer solstice. The fish-eye photo from Figure 8 indicates that the painting receives direct sun year-round for about 55% of the morning hours. The horizontal ceiling of the portico protects the paintings from the sun for about an hour everyday of the year. Figure 9 shows the amount of sunlight penetrating this gallery on March 5 at 9:00 AM. Illuminance levels measured over the paintings at this time reached values of 2,400 lux, which is about 12 times higher than the recommended IES standards for moderately light susceptible display materials. Every night each of the oil paintings of this gallery are covered with boards to protect them from the morning sun and UV radiation. Every morning, the boards are removed from the painting few minutes before the museum open its doors for visitors (10:00 AM).

Figure 8: Fish eye photo taken from painting’s Figure 9: Entrance lobby at the Amon

viewpoint with sun path diagram, at Amon Carter Museum with direct sunlight,

Carter Museum’s east gallery (main lobby). March 5, 2004, at 9:00 AM.

South-facing gallery (2nd floor):

This gallery is the only one on the second floor that receives natural light directly from a side light window. The gallery is located right over the South entrance to the museum.

The gallery displays mainly oil paintings and metal sculptures (Figure 10). The 210-ft2- window area (19.5 m2) has a five-feet (1.5 m) external horizontal overhang, and the visible transmittance (Tvis) of the glass is 12%. The window wall ratio (wwr) of the gallery is 58%, and the window floor ratio (wfr) is 28%.

Direct sunlight inundates the gallery all day throughout the year. Fish eye views taken from the painting’s viewpoint show that it receives direct sunlight between 2:00 and 4:00 PM from November to January (Figures 11 and 12). Illuminance measurements taken over the painting under direct sun reached up to 2,200 lux, which is 11 times higher than the IES recommended standards for oil paintings. The total illuminance-hours during these two hours of sunlight over the painting is around 404,800 lux-hours, when added the illuminance-hours over the painting during the rest of daylight hours 700,000 lux-hours, the total over exceeds the maximum annual exposure to light recommended by IES for oil paintings. Figure 12 also shows that the horizontal overhang blocks sunlight few hours around wintertime, but does not shade enough the window to protect the painting. Paintings in this gallery are exposed to daylight at all times without any device that could help to reduce the illuminance levels over light susceptible artwork.

Figure 11: ECOTECT’s stereographic diagram taken from painting’s viewpoint, at Amon Carter Museum’s south gallery.

Figure 10: South-gallery at the Amon Carter Museum with direct sun, November 28, 2003, at 3:40 PM.

04

Figure 12: Fish eye photo taken from painting’s viewpoint with sun path diagram, at Amon Carter Museum’s south gallery.

Figures 13 and 14 show the illuminance levels simulated with the Desktop RADIANCE lighting program in the south gallery on November 28, at 3:40 PM. Simulated illuminance levels were calibrated with the measured illuminance levels during site visits. Lighting simulations were done at different times during the day to evaluate the sunlight patterns in the gallery. Results from these simulations showed that the display areas over the walls

receive direct sun in the morning (west wall) and afternoon (east wall) for about two hours around winter solstice.

Figure 14: Radiance model of Amon Carter Museum’s south gallery on November 28 at 3:40 PM; false color image with illuminance levels (bottom).

Thickness K K cm w/mq K w/mq K External structural walls external plaster 1,5 Poroton Aktuell with insulation plaster 49 0,22 internal plaster 1,5 e x t Roof tiles 1 e air chamber 5 r waterproof layer 0,5 0,26 n Extrude polystirene insulation 12 a precompressed wood slab 1,5 i Pavement ventilated air chamber 50 solaio pignate e travetti precompressi 15 massetto 5 0,46 f Extrude polystirene insulation 5 a radiant pavement 10 c ceramic 1 e s Windows abete wood frame 6 1,67 low-e glass in layer 2 (solar gain) 0,4 argon gas chamber 1,2 1,1 1,5 glass 0,4 Tab. 1 — Characteristics of the external surfaces . The space heating energy demand

The house has been simulated with the DEROB LTH (Dynamic Energy Response Of Buildings) version 00.04, developed by the Swedish Department of Building Science belonging to the Lund Institute of Technology. Natural ventilation has been considered

2609 kWh/y for heating (29 kWh/m2y)

Fig. 8 — Model developed by DEROB LTH simulation programme

during the whole year.

The results indicate that the volume A will re and 2812 kWh/y for cooling (30 kWh/m2y). This is a lower demand compared to the heating demand of a typical Italian residential building.

4.2 The heating systems (solar and biomass)

Since the energy consumption for heating is low, a great part of it could be covered by a solar heating system. Therefore two solar heating systems have been designed: a water solar system with solar collectors to cover a great part of the heating demand and the DHW needs (Costruzioni Solari s. r.l.[16]) and an air solar system (Solarwall[17]) to preheat the inlet air during the winter sunny days. The

Fig. 9 — Conventional solar system winter behaviour

solar system will heat the house through a radiant pavement system at low temperature. The whole solar system is integrated with a wood stove to cover the complete heating demand during the coldest period.

The water solar system

Fig. 10- Solar water system for space and domestic water heating scheme

Six solar thermal collectors of 1,9 sq meters each and one boiler of 700 litres for the space heating system are located in the south wall as reported in figure 9. The solar system scheme is reported in figure 10. This system should cover from 64% to 100% of the heating demand. In order to increase this percentage, a solar air system has been designed.

Days/ month

Month

Days /month

Average daily radiation in the sloped surface

Average

system

efficiency

(Qa) Daily

average

thermal

energy

available/ sq

m

(Qa) Monthly thermal energy available/ sq m

Monthly

thermal

energy

available

(Ea) monthly

energy

demand

Surplus/integ

ration

% solar fraction

kWh/m2 day

kWh / m2 day

kWh / m2 month

kWh/

month

kWh/

month

kWh/

month

%

31

January

31

3,14

0,40

1,26

39

443

600

— 157

74%

28

February

28

3,42

0,40

1,37

38

437

542

— 105

81%

31

March

31

3,81

0,45

1,72

53

606

600

6

101%

30

April

0

0

0,50

0

0

0

31

May

0

0

0,50

0

0

0

30

June

0

0

0,50

0

0

0

31

July

0

0

0,50

0

0

0

31

August

0

0

0,50

0

0

0

30

September

0

0

0,50

0

0

0

31

October

0

0

0,50

0

0

0

30

November

30

3,04

0,45

1,37

41

467

581

— 113

80%

31

December

31

2,73

0,40

1,09

34

385

600

— 215

64%

365

TOTAL

151

3,23

0,47

1,36

205

2.338

2.923

80%

Table 2 — Heat production and the coverage (in %) of the solar system.