Only a part of the sun radiation which penetrates the atmosphere reaches the earth surface directly. Another part of the solar radiation is scattered by the atmosphere and reaches the earth surface indirectly as diffuse radiation.

In areas of a temperate climate, diffuse daylight manifests itself most clearly on a cloudy day in winter. To simulate diffuse daylight the CIE-overcast sky is used, of which the luminance L varies with the vertical angle a. In formula:

TOC o "1-5" h z L = ((1 + 2 cosa)/3) LZenith. (1)

where LZenith is the luminance of the highest part of the sky.

The illuminance in a free field situation (Ehor, ff) is:

Ehor, ff LZenith 7я / 9~ 2.44 LZenith (2)

The daylight factor (df) for a point P inside a building is defined as

df = Ep/ Ehor. ff (3)

where Ep is the illuminance at point P in the building.

The direction of the direct sunlight is defined by two angles, the azimuth (a) and the height (h). The formula for the height of the sun is:

h = arcsin ( sin 9.sin d — cos 9.cos d. cos u) (4)

where ф = the degree of latitude of the place on earth; d = the angle of the declination; u = the hour angle, the angle corresponding with the daily rotation of the earth.

The formula for the azimuth is:

a = arcsin {( cos d. sin u)/cos h} (5)

The azimuth and the height of the sun vary during the day and depend on the place on earth. The computer program Radiance can be adjusted for every location.

Various skylight models had been prepared with different geometry and transparence surface characteristics. The efficiency of the built skylights had been determined in the horizontal "reference” plane that is underneath the horizontal uniform square opening of the skylights — this is the same "theoretical” surface, where the skylight is connected to the lighting well of the daylighting system.

Ф *extemal is the luminous flux that would be absorbed by the 6x6cm reference plane without the presence of the built skylight. The external horizontal illumination had been kept at a constant Eexternal = 428 lx value. The illuminance values had been determined in the reference plane with the help of a measuring head in the grid-points of a uniform 1x1cm measurement grid. This 1x1cm measurement grid (x’;y’) had been revolved by 45° compared to the 6x6cm reference plane of the skylight. The measurement grid has i = 1 ^ 7 measurement points in both (x’) and (y’) directions, that is 49 measurement points altogether. The illuminance values E(x’^/j) — measured in the (x’;y’) grid-points — multiplied by the corresponding area of the measurement points A{x’i;y’j_) provides the luminous flux 0{x’j;y’j) absorbed by the reference plane. The rjskyligM efficiency of the investigated skylight had been calculated as the ratio of the 0extema, luminous flux absorbed by the unobstructed horizontal plane and of the 0reference plana luminous flux

absorbed by reference plane beneath the investigated skylight model using the following equation:

The efficiency of the illustrated "Arched-Shed” skylight geometry with transparent, simple glazing is nArchedShed = 0.41. The investigated skylight characteristics and the corresponding efficiency values can be found in the two-pages attachment.

A simple steady state model of the TCA has been created in EES (Klein and Alvarado, 2003). The model is built in four parts:

• Heat exchanger for condenser/evaporator

• Pressure drop calculation for mass transport

• Effective AT between vessels

• Heat exchanger for reactor

The heat exchangers are modelled as a simple counter-flow heat exchanger with fixed UA — value, using the effectiveness relationship of Equ. 3 (Incropera and DeWitt, 1996). Cmin is the lower of the two heat capacitance rates, and Thi/Tci are the hot and cold inlet temperatures respectively. The effectiveness is a function of the UA-value and capacitance rates.

Qt є Cmjn (Thi — Tci) Equ. 3

In the tested prototypes, the heat exchangers were spirally wound smooth tubes. On the outside of the tubes there is condensation, evaporation, desorption or absorption depending on the vessel concerned. All of these processes are assumed to take place at essentially constant temperature over the heat exchanger. This is implemented in the model by making the heat capacitance rate on the outside of the tubes 1000 times greater than that on the inside. The heat transfer coefficients used in the model are given in Table

1.

Where:

f Friction factor

L Length between evaporation and condensation

D Hydraulic diameter

p Density of steam

vvap Velocity of the water vapour

v Viscosity of the vapour

The velocity of the vapour is directly proportional to the rate of evaporation or condensation and thus the heat transfer rate, assuming no heat losses. Using this fact and putting all factors relating to the geometry into one coefficient CAP, the pressure drop during discharge becomes:

Properties for the vapour are taken for the temperature in the evaporator (Te), where the mass transport is most critical due to the low pressure.

The effective temperature difference between the reactor and the evaporator (ATeff, sr) is calculated in a number of stages. First the mass fraction of LiCl required for saturated solutions at the theoretical reactor temperature (Tsr*) is calculated (Equ. 7) using the relationship derived for the solubility line (Conde, 2004). Next the vapour pressure for this saturated solution at Tsr* is calculated using another relationship derived by Conde (Equ. 8), as is that for water in the evaporator (Equ. 9). These two pressures are related in Equ. 10 by the pressure drop due to the mass transport (APcooi) as calculated by Equ. 6. The effective temperature in the reactor (Tsr) is different to the theoretical temperature by ATsubcool (Equ. 11), a measure of the imperfection in the absorption process. The effective temperature difference (lift) is then calculated by Equ. 12.

It is assumed that a certain fraction (Fchem) of the heat transferred to the TCA during is stored chemically in the TCA, and that during discharge this is released (Equ. 13). This implicitly assumes zero losses from the TCA and neglects losses due to the cooling down of the vapour between the reactor and the condenser during charge and of heating the vapour between evaporator and reactor during discharge. Fchem is set to 0.3 in the model. The temperatures for the charging process are calculated in a similar manner.

Comparison with Measured Data

A prototype unit was tested both by the producing company, ClimateWell (then called Solsam), and by Vattenfall AB (Setterwall and Bales, 2003). The tests were carried out for both charging and discharging, using constant inlet temperatures to the condenser (Tcii) and reactor (Tsrii) respectively. Figure 5 shows the temperatures relevant for these measurements.

The required charging temperature (Tmr, i) is plotted against the charging rate (Q’mr) in Figure 6 a), and the chiller supply temperature (Te, o) is plotted against the cooling rate (Q’e) in Figure 6 b). ATsubcooi was adjusted to best fit the measured data, one value each for charging and discharging, whilst CAP was estimated to be 20 based on the geometry of the prototype.

a) b)

Figure 6. Comparison of model output (solid lines) with measured data (symbols) for three different boundary conditions, for a) charging, and b) discharging (cooling).

A value of 15°C was required for ATsubcool in order to give good agreement with measured data for discharging. This is higher than can be achieved in low-temperature absorption chillers (Foley et al., 2000), and shows that there is a large potential for improvement in
heat exchanger design and thus available temperature lift. A figure of 3-5°C should be possible, and work is in progress to improve the heat exchanger design. A value of 7.5°C was required for ATsubcooi, in order to fit the measured data for charging. This correction term is not usually required for the generator in absorption chillers and the relatively large value could be due to the fact that crystallisation takes place on the tubes. Alternatively the equations derived by Conde could be inaccurate in this temperature range. Further measurements of vapour pressure above saturated solutions of LiCl are required for temperatures of 60 — 100°C solution temperature. These could not be made at ClimateWell due to limitations of their measurement equipment.

Conclusions

The principles of operation of the TCA have been described in general terms. The process offers some advantages over other heat driven cooling process:

• High energy density storage in the solid crystals, with 180 kWh/m3 based on the volume of the main reactor and condenser vessels. This is of the same order as for prototype thermal stores using adsorption of water with silica gel.

• Good heat and mass transfer, as this occurs with solution. The heat transfer characteristics are similar to those for absorption chillers, but significantly better than for adsorption chillers.

• Constant operating conditions, with constant temperature difference between the reactor and condenser for a given solution temperature.

The TCA has desorption temperatures of below 100°C for ambient temperatures below 40°C, which is relatively low. The temperature lift depends on the cooling rate supplied and varies from 15°C for design cooling rate of 5 kW per TCA unit and 30°C inlet temperature to the reactor, to 20°C for a cooling rate of 2.5 kW.

The TCA has so far been tested only with LiCl as the salt. Theoretically it is possible with other salts, but these have not been tested. Physical properties of LiCl have been found in the literature, including equations for the most important properties including solubility line and vapour pressure. Measurements made by ClimateWell agree well with these relationships for temperatures up to 60°C, but further measurements are required above this point to check whether the relationships from literature are valid for these temperatures, relevant for charging conditions in the TCA.

A simplified steady state model of the TCA has been made based on the physical properties of LiCl as defined in the literature and a heat exchanger model with fixed UA — value. This model needs to be refined as follows:

• Calculation of the fraction of energy stored chemically (Fchem) based on the physical properties of LiCl.

• Calculation of heat losses due to changes in temperature of the vapour between reactor and condenser/evaporator.

• Estimation of losses from the TCA vessels to ambient.

This model has shown that the absorber heat exchanger works significantly worse than absorbers in equivalent low-temperature absorption chillers, and that this is a critical component for future development.

The model of the whole system is created in the TRNSYS program (Klein, 2000) but the mathematical model of the refrigeration subsystem is written in the Engineering Equation Solver (EES) Program (Klein, 2002). The TRNSYS calls the refrigeration sub-model, which is built in EES, by component type 66. The component type 66 calls an EES file, passes it information and receives outputs from EES using the dynamic data exchange method and text files to exchange the information and commands. Properties of the refrigerant (butane) are taken from NIST reference database REFPROP (McLinden, 1998). The weather data of Bangkok, Thailand is used and it is generated from METEONORM version 4.10
(Remund, 2001). The analysis of the solar-driven ejector refrigeration system was performed assuming the following conditions:

Solar collector subsystem

Flat plate, double-glazed solar collectors are used (TRNSYS type 1d). The intercept efficiency (FR(xa)e) is set at 0.9. A 2 m3 storage tank (TRNSYS type 60) is used in the standard case when studied the effect of the solar collector area. An auxiliary heater (TRNSYS type 6) is placed between the storage tank and the refrigeration subsystem for reliability of the system when lacking of solar radiation. The outlet temperature from the auxiliary heater is assumed to be 10 K above the generating temperature. A pump in the solar collector subsystem represented by a TRNSYS component type 3b. Water is used as the heating medium in the solar collector subsystem. A signal from the controller (TRNSYS type 2b) controls the operation of the circulation pump. The pump works only when the output temperature of the solar collector is higher than the input temperature 10 K.

Refrigeration subsystem

Butane is used as the refrigerant for the refrigeration subsystem. The condensing temperature and the evaporation temperature are 30°C and 10°C respectively. The pump in the ejector subsystem is assumed to have an efficiency of 50%. The total efficiency of the ejector is assumed to be 70%. This system is designed for the minimum generating temperature of 80°C and the maximum generating temperature of 150°C. The model is written in EES as mentioned above.

Cooling load

The load is modeled by using TRNSYS component type 16 with the energy rate control mode. The building volume is assumed to be 150 m3 (5m x 10m x 3m), with 9 m2 windows area (3 m2 facing north and 6 m2 facing east). The floor is built from a heavy weight concrete block and the ceiling is built from a heavy weight concrete. The infiltration rate is assumed to be 0.5. The cooling load starts at 7:00 and stops at 17:00. The building is assumed to be an office, which has 5 workers and 0.28 kW of internal load from electrical equipment. The indoor temperature is set at 25°C.

Liliana O. Beltran, Ph. D., Texas A&M University

Abstract: This paper presents the assessment of the lighting conditions of selected exhibit areas in three museums located at the same site, the Dallas-Fort Worth Cultural District. These museums are the Modern Art Museum by Tadao Ando (2002), the Kimbell Art Museum by Louis Kahn (1972), and the Amon Carter Museum by Philip Johnson (1961, 2001). Each of the museums presents different lighting conditions to exhibit their art collection. This study focuses on specific galleries that include daylight as a source of illumination. Each selected gallery is examined; assessed on the site; simulated using state-of-the-art lighting tools; and evaluated according to good lighting practice: light exposure, glare, and ultraviolet radiation.

Introduction

Light in museums is necessary to enhance and view museum objects; but at the same time, if not properly controlled, light can harm and reduce the life of the museum artifacts. To improve the quality of illumination in museum galleries, daylight has been reintroduced after many windowless galleries were built. Daylighting compared to most electric light sources has more potential to harm susceptible objects, and represents a challenge for designers to succeed in controlling it. Sunlight can create serious problems in the conservation of museum objects, when window systems are not carefully designed to intercept sunrays at all times.

This paper presents the evaluation of three museums that have many similarities as well as differences, and the reflections about the design of daylighting systems in museums. These museums, the Modern Art Museum, Kimbell Art Museum, and Amon Carter Museum, are located in the Fort Worth Cultural District, next to each other (Figure 1). These museums that include in their collection valuable art work, were designed by three well-known architects, who have used a variety of daylighting systems to illuminate the museum galleries to best accommodate the exhibit’s lighting requirement.

The contribution of the solar facade to reduce energy consumption derived from the ne­cessity of space heating in a building, has been quantified by means of the solar fraction delivered by the facade. Heating load has been calculated in base to degree-days method (annual base 15: 573 for Barcelona and 1377 for Geneve), considering a typical residential unit of 90 m2 of habitable surface and a volume equal to 225m3. 2m2 of integrated collector (as described in Table 3) is applied in the south wall. The global heat loss coefficient includes the losses by ventilation, considering an air renovation rate equal to 0.5 renovation by hour. Table 7 shows the heating load for different types of good insulated dwellings in Barcelona climate and Table 8 represents the heating load for three insulation levels in Geneve climate.

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(a)

Figure 7: Monthly solar fraction and solar efficiency of the facade applied to space heating for both climatic conditions, a — From January to May, b — From October to December

From Figure 7 it is observed that in May and October, the facade is able to deliver the total heating load in the majority of the cases, the rest of the months the solar fraction is lower. For Barcelona climate, the facade is more efficient. Table 9 shows the values of the useful energy (QLOAD) delivered by the facade for both climatic conditions.

For Barcelona climate, it may be obtained average annual values of solar fraction from 32% to 20% depending on the type of construction. For Geneve climate, annual values are about 9%, in this case, larger surfaces for collecting solar energy would be necessary.

Conclusions

Facades play an important role in the thermal behaviour of the buildings or dwellings at which they are addressed. Large glazed areas implemented in double skin facades, allow the combination with partially inner opaque layers. This paper has shown the possibility of using these areas to collect and accumulate significant quantities of solar energy in water tanks placed at the facade. A 2 m2 integrated solar collector implemented as part of the facade accumulating 100 water liters, permits to get annual values of solar fraction of around 41% for a Mediterranean climate (Barcelona) and 31% for a central Europe climate (Geneve) when it is applied to domestic hot water assistance. For space heating applications, the annual solar fraction for Barcelona climate is around 20% whereas for Geneve climate it is
about 9% (depending on the degree of insulation of the construction). For Geneve climatic conditions it would be necessary larger collection areas. Numerical results not shown here, have demonstrated that annual performance does not vary outstandingly with larger storage volumes, since these variation is asymptotic from 50l/m2 [9], [2].

Annual solar efficiencies for both applications are similar, for Barcelona climate it is around 48% and for Geneve, it is about 44%. The geometric configuration and material properties of the prototype analyzed in this paper, was previously optimised by means of a permanent analysis using code. The configuration with lower heat loss coefficient was chosen.

Numerical code AGLA has demonstrated its strength to simulate advanced facades inclu­ding the special features described in this paper.

Acknowledgements

This work has been supported in part by the European Commission under the Fifth Framework Programme, Thematic Programme: Energy, Environmentand Sustainable De­velopment FP5-EESD, Project CRAFT-1999-70967.

[1]

During the design process, it took little to convince the Council to utilise the new and different technologies and ideas. After justification of the benefits, as well as invariably the payback, they were able to see the value in these decisions and agreed with the choices made. This was particularly applicable to the WaterFurnace Geothermal system, which has a 4.03 year payback, and a 37% saving on life cycle costing (refer Tables 1 & 2).

Table 1 — Simple Payback Analysis

The principle of the Geothermal system is a simple heat exchange system. An underground loop is filled with water and sunk down into the ground. In this case, 38 bore holes, some 94 metres deep, sunk into the ground external to the building and now hidden under landscaping, provide the necessary ground loop to run the system, while the pumps are housed in an existing area which became an external pump room.

As the earth’s temperature is a constant 15-20 degrees C once past a depth of approximately 2m, the water in the pipe exchanges heat with the ground temperature depending on the current conditions. It is via the pipe that heat is dissipated when cooling in summer, and heat is absorbed when warming the building in winter. The energy then required to bring the temperature up or down to the desired range is significantly reduced. Thus in winter, the heating system gets say a 145 degree boost if the outside temperature is only 0 C whereas the underground temperature is already at 15. Similarly, in summer the cooling system is able to unload the hot air into 15 being the underground temperature, instead of sat 35 — being the outside temperature. On average, the energy saving amounts to a 400% efficiency — ie for every kW pumped into the system, there is a 4kW return.

Issues of accessibility and security were also part of the brief, and both have been resolved to the mutual benefit of all users. External disabled access has now been provided, and all doors into publicly accessible areas are wide enough to accommodate wheelchair usage. Disabled toilets have been included with the facilities, and a small lift provides access to the upper floor.

Security between public and non-public areas was high priority, but without the sense of ‘loosing’ the rest of the building, and the benefits of the voids. One solution has included a separation between the two areas being divided with a glass display cabinet, allowing light to filter through without compromise to security.

The concept of integrating the parabolic collectors into the cycle is illustrated in figure

4.

Because the temperature level of the solar steam is high enough, two bypasses with solar preheaters are integrated into the high-pressure and low pressure prehea­ter section of the water steam cycle.

In the solar preheater 1 the superheated steam gets condensated via heat transfer to the feedwater, which is switched into the bypass by a control valve. This heat flow is permanently varying with a maximum at noon caused by the changing irradiation in the course of the day. The condensate leaving solar preheater 1 is still at a tempera­ture level to preheat a part of the feedwater in the low pressure preheating section (solar preheater 2). The main part of the feedwater is heated in the preheaters 1 to 4 by extraction steam conventionally.

Since the solar generated steam preheats a part of the boiler feedwater, the steam consumption of the conventional preheaters 1 to 7 is reduced.

The steam remaining in the turbine can be used for an additional dynamic power generation.

A special control strategy consisting of 2 control loops has been developed that al­lows the addition of the external heat into the preheating cycle without affecting the feed-water temperature at the boiler inlet (251°C) and therefore without affecting the complex boiler control. The temperature of the feedwater through the preheaters 6 and 7 is controlled by steam valves that are installed in the steam extraction lines between the turbine and the preheaters. In order to avoid any mixing of “cold” bypass feedwater with the mainstream in the morning and evening, it gets discarded by a control valve as long as it is below the set point temperatures of 251°C in the HP sec­tion and 150°C in the LP section.

Results

In figure 5 the results based on the average irradiation in July are illustrated.

It can be seen that the water at the outlet of the parabolic trough (PT) gets warmed up in the early morning (t=5h), is then vaporized (t=6h) and finally superheated to a controlled constant temperature of 380°C. It is obvious that the PT outlet temperature follows quite closely the irradiation curve. In parallel to the irradiance the feedwater mass flow through the parabolic trough increases till noon and then decreases.

Due to mass release effects during the vaporization phase in the morning the mass flow characteristic through the absorber tubes shows a peak. Dynamic effects like these in combination with long dead times resulting from long absorber tubes and low flow velocities complicate the development of reliable automation concepts for the water mass flow through the absorber tubes in the parabolic trough collectors [9]. Therefore a pilot control concept based on measured irradiance data to pre-adjust

the water mass flow has been applied. Figure 6 illustrates the resulting mass flow characteristics in the high pressure preheating section during the course of a day in July.

As can be seen, up to 179 kg/s of feedwater can be guided into the bypass to be preheated by the solar generated steam. In consequence, the mainstream feedwater mass flow decreases from 320 kg/s to 141 kg/s. Since the reduced feedwater mass flow needs less steam the controller closes down the extraction steam valves as it is shown in figure 7. The processes appearing in the low pressure preheating section are very similar. Closing the steam turbine extraction lines leads to a significant in­crease of the steam power plant’s power output as shown in figure 8. The power in­crease is following the solar irradiance characteristic nearly without any time delay — this is an essential result of the investigations carried out.

The combination of integrating an external heat source into the cycle and closing the extraction steam lines leads to a rise of the steam power plant output from 393 MW up to 410.5 MW at noon in July. This corresponds to a power increase of approxi­mately 4.5 percent. Compared to these results, the potential power increases in the other months are lower, for example in January the maximum additional power rises about 6.7 MW.

On the basis of this simulation data an efficiency factor can be calculated, defined as a ratio of the parabolic trough thermal power output integrated into the cycle and the achievable electrical power output increase at the steam turbine’s generator. This efficiency factor is about 28 % at noon.

In addiction to the use of average irradiance data for the simulation a single day in July showing strongly varying irradiation caused by clouds was selected as a test case. Because the irradiance is temporarily too low to preheat the feedwater in the solar preheater 1 to 251°C, the feedwater is not led back to the mainstream but dis-

carded. In these periods no additional power is generated in the steam turbine. Hence significant output losses are to be noticed in the example simulated.

However, the results favourably show, that even fast changes in solar irradiation can be managed with simple control strategies.

Conclusion

The integration of external solar generated steam into the water steam cycle is a promising concept to promote renewable energy and as a consequence to reduce the CO2 emissions. The simulation results, based on meteorologic data of a location in Italy, show a power increase following the solar irradiance characteristic nearly without any time delay. This can be managed even with simple control strategies. Since the peak of energy consumption is to be found at noon, the additionally pro­duced power matches very well the electricity demand of industrial countries.

The efficiency factor defined as the ratio of the parabolic trough thermal power output integrated into the cycle and the achievable electrical power increase at the steam turbine’s generator amounts 28% maximum.

• The solar radiation field:

In the estimation of the solar radiation we used the models reported in the "Atlas solaire d’Algerie” [2]. The global solar radiation is the sum of a direct and a diffuse radiation. A critical report on the capderou work showed that if the direct radiation is underestimated the diffuse radiation for a mean day is quite surestimated compared to measured values. All the equations used to estimate the direct and diffuse radiations for any site and given day can be found in reference [1].

The energy balance over the flat solar collector leads to the following system of equations solved by the 4th order Runge-Kutta method.

(Tf and Tp are the fluid and the collector absorber temperatures respectively)

• The ejector flow loop:

• entrainment ratio ra :

The entrainment ratio is equal to the ratio of the secondary and the driving mass flowrates.

with m’ : driving mass flow rate and m’’ : secondary mass flow rate The optimal ratio for any working fluid is given by the following equation [2] :

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The installation of the plant and the first two years of scientific monitoring were funded by the European Commission (DG TREN) and the federal state of Baden-Wurttemberg. In 2004 the monitoring and some technical improvements are funded by the German Government as one of the demonstrations projects within the IEA Task 25 "Solar air conditioning of buildings”.

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425 — 430, Shanghai, China, September 2002.

/8/ U. Eicker. Solare Technologien fur Gebaude, B. G. Teubner Verlag, Stuttgart/ Leipzig/ Wiesbaden, 2001. /9/ U. Franzke, G. Heinrich (Hrsg.). Sorptionsgestutzte Klimatisierung: Entfeuchtung und DEC in der Klima — Kalte — Technik. C. F. Muller Verlag, Heidelberg, 1997.

/10/ Hans-Martin Henning (Ed.). Solar — Assisted Air — Conditioning in Buildings. Springer Wien New York 2004.

/11/ C. Hindenburg, H.-M. Henning, G. Schmitz. Einsatz von Solarluftkollektoren in sorptionsgestutzten Klimatisierungssystemen, Achtes Symposium thermische Solarenergie; OTTI, Staffelstein; Mai 1998.

/12/ C. Hindenburg, L. Schnabel, T. Geucke, M. Motta. First thermally solar autonomous air conditioning system in Germany — Simulations, operation experience and economic aspects. Proceedings of the ISES Conference, Gothenborg/Sweden, June 2003.

/13/ Deutsches Institut fur Normung DIN 1946 Teil 2: Raumlufttechnik/ Gesundheitstechnische Anforderungen (VDI-Luftungsregeln), Beuth — Verlag 1994.

/14/ J. Pfafferott, S. Herkel, A. Wagner. Sommer 2003: Mussen unsere Burogebaude klimatisiert werden? HLH 55 (2004)

/15/ L. Schnabel. Betriebsanalyse und Simulation einer offenen sorptionsgestutzten Klimatisierungsanlage. Diplomarbeit. Technische Universitat Berlin. 12/ 2003.