Category Archives: Sonar-Collecttors

Model Measurements in the Artificial Sky

Figure 4: Artificial Sky at the T. U.B.

Figure 5: Cross-Section of the Artificial Sky

The most convenient, reliable and precise solution to describe and follow the illuminance distribution and efficiency of the skylight is to perform scaled model measurements in the artificial sky. The main advantage of these physical model measurement it is possible to follow and analyse all the necessary and investigated material and geometry properties and characteristics — even without knowing the exact mathematical and physical background of the events. In case of the artificial sky the Input can be precisely determined — since any standard CIE sky condition can be set-up. With the help of the scaled, physical model of the daylighting system — the Output can also be measured enabling to analyse, compare and evaluate all the investigated illumination characteristics of actual skylight. The accuracy of this method meets all practical requirements. The sky itself is a 6 m diameter hemisphere, illuminated around its parameter under its horizontal plane. All required standard sky conditions — i. e. CIE Overcast Sky — can be achieved and maintained throughout the measurements. The daylighting model can be placed in the middle of the sky, and inside the model all the illumination values can be measured and the efficiency of the system can be precisely analyzed.

Existing calculation method

to determine light distribution and illuminance efficiency

1.

2.

3.

4.

5.

6.

Transmis­sion of trans­parent structures

Clear or diffused transpa­rent

structures

Obstruction of trans­parent structures

Reflection of trans­parent structures

Shiny or matte light — guiding surfaces

Geometry of skylight

I. Grunn Method

Yes

No

No

No

No

Partly

II. Daniluk Method

Yes

No

Yes

Yes

No

Partly

III. C. I.E.16. Pub­lication, "Skylights”

Yes

No

No

No

No

Partly

IV. C. I.E.16. Publ., "Monitor skylights”

Yes

No

No

No

No

Partly

V. C. I.E.16. Publ., "Shed skylights”

Yes

No

No

No

No

Partly

VI. Lumen Micro 7.0 computer software

Yes

Yes

No

No

No

Partly

VII. TTI TS-A5 Dim. "Dome skylights”

Yes

No

No

Yes

No

Partly

VIII. TTI TS-A5 Effi­ciency of skylights

Yes

No

Yes

No

No

Partly

IX. S. Birch, I. Frame, Daylight

Yes

No

No

No

No

Partly

X. B. R. S. Method

Yes

No

No

No

No

Partly

XI. Lightscape computer software

Yes

Yes

No

No

No

Partly

XII. Skylight Dimen­sioning Method

Yes

Yes

Yes

No

No

Partly

Table 1: Existing Daylighting Calculation and Dimensioning Methods

Existing Daylighting Calculation Methods

The table above (Table 1.) indicates some of the existing investigating methods, and it also lists the properties, taken into consideration to determine the illuminance distribution in the space bellow. None of the listed existing methods or computer software consider all the properties of the skylights, which are playing an important and basic role in the light distribution and efficiency.

These methods are not able to differentiate between shiny or matte surfaces and they are not able to handle all the geometrical properties of a certain light modulating structure. At the calculation of diffused surfaces the transmission values are taken into consideration on a very base level, because these formulae are not including inclination angles of surfaces, which can play very important and determining role in some case.

These facts are proving that the existing calculation methods are all neglecting some important features, this is leading to the necessity to handle the problem using a new and different system of investigation. The model measurements in the artificial sky are able to provide a more exact and global answer to this question.

Working Principles of the TCA

The thermochemical accumulator (TCA) is an absorption process that uses a working pair, not only in the liquid, vapour and solution phases but also with solid sorbent (Olsson et al., 2000). This makes it a three-phase system, with significantly different properties from the traditional absorption processes, where there are only two phases: either solution + vapour or solid + vapour. Figure 2 shows the schematic of a single TCA unit, which is similar in
principle to that of Figure 1. In a practical unit the vessels are evacuated and the solution is pumped over a heat exchanger to increase the wetted area and improve heat transfer.

During desorption in the reactor, the solution is saturated and further desorption at the heat exchanger results in the formation of solid crystals that fall under gravity into the vessel. Here they are prevented from following the solution into the pump by a sieve, thus forming a form of slurry in the bottom of the vessel.

This gives the TCA the following characteristics:

Figure 2. Schematic of a single unit thermochemical accumulator.

High energy density storage in the solid crystals.

• Good heat and mass transfer, as this occurs with solution.

• Constant operating conditions, with constant ATequ for a given solution temperature.

For discharging, where the process is reversed, saturated solution is pumped over the heat exchanger in the reactor where it absorbs the vapour evaporated in the evaporator using the cooling load. The solution becomes unsaturated on the heat exchanger, but when it falls into the vessel it has to pass through the slurry of crystals, where some of the crystals are dissolved to make the solution fully saturated again. In this way the solution is always saturated and the net result is a dissolving of the crystals into saturated solution.

LiCl Porperties

The first TCA units have been built using water/LiCl as the active pair. The physical properties of this pair have been summarised in the literature (Conde, 2004) and empirical equations have been created for them based on data from a large number of studies over the last 100 years. The solubility line for LiCl can be seen in Figure 3 a), where it is readily apparent that there are several different hydrates for LiCl. However, for the operating range of the TCA, the solution is generally operating at temperatures of 20-50°C for discharge and 65-95°C for charge, all of which are within the monohydrate range for saturated solution. The figure shows that the mass fraction for saturated solution is a function of the solution temperature, and thus Equ. 1 is simplified to Equ. 2 for the TCA. This in practice means that ATequ is constant for a given set of boundary conditions resulting in constant operating conditions during charging/discharging.

ATequ ~ fsat (Tsol) Equ. 2

The equations for vapour pressure derived by Conde were used to create a Duhring chart for LiCl, Figure 3 b). This shows that the maximum value for ATequ is for the saturated solution, and that for the operating conditions of the TCA with an ambient temperature of 35°C, ATequ is 37°C for discharging (comfort cooling), and 53°C for charging.

SHAPE * MERGEFORMAT

Figure 3. Data for LiCI: a) solubility line (Conde, 2004); b) water vapour pressure above the solution for varying mass fractions of sorbent and solution temperature (Tso). Measurements made by ClimateWell using the solution used in the TCA are shown as filled squares.

Figure 4. Relationship of ATequ to the

saturated solution temperature.

Figure 4 shows the relationship of ATequ to the temperature of the saturated solution. ClimateWell have made their own measurements at different times with the mixture of LiCl that they use in the TCA. The lower line shows the correlation ClimateWell use in their control system, whereas the filled squares represent data for the latest measurements. The data from 2004 agree well with Conde’s equations at higher solution temperatures, but deviate somewhat at lower solution temperatures, as can also be seen in Figure 3 b).

A SOLAR-DRIVEN EJECTOR REFRIGERATION SYSTEM

A solar-driven ejector refrigeration system consists of two main sub-systems: a solar collector sub-system and a refrigeration sub-system. The major components in the system include solar collectors, a storage tank, an auxiliary heater, an ejector, a condenser, a regenerator, an evaporator, an expansion device and pumps.

The solar collector subsystem

This subsystem consists of solar collectors, a storage tank and an auxiliary heater. Solar radiation is converted to heat by the solar collectors; the heat is transferred to the heat supply medium (water) in the solar collector, then it is stored in a storage tank before being supplied to the refrigeration subsystem in the generator. The auxiliary heater is placed between the storage tank and the generator. The lowest generating temperature of the refrigeration subsystem is set at 80°C. The temperature difference between the heat source fluid (water from the storage tank) and the refrigerant in the generator is assumed
at 10 K. If the temperature of the heating medium is lower than 90°C the auxiliary heater will start and the working medium is heated until it reaches the set point.

Qu

AJ

FrUl (T — Ta)

(1)

шСр (T0 — T)

AJ

= FrI (та) e

In the model of the flat plate solar collector in TRNSYS, the solar collector efficiency is calculated from the heat balance in the flat plate solar collector by the Hottel-Whillier-Bliss equation. It is basically defined in the form of the average Bliss coefficient (FR(xa)e) and the heat loss coefficient (FRUL).

An ejector refrigeration subsystem

The main energy supply to this subsystem is heat, but a small amount of electricity (supplied to the pump) is required to circulate working fluid inside the system. In the refrigeration subsystem, the high velocity vapor stream (from the generator) goes through a converging-diverging nozzle in the ejector resulting in the vapor being sucked from the low temperature evaporator. Suction occurs, as the pressure is low at the narrowest section of the ejector. The stream from the evaporator reaches subsonic velocity. In a mixing zone at the end of the converging section, the two streams are mixed. After mixing, the combined stream becomes a transient supersonic stream and the velocity of the combined fluid must be high enough to increase the pressure after deceleration in the diffuser to a suitable condensing pressure. The vapor from the ejector goes to the condenser, condenses and heat is rejected to the environment. After the condenser, part of the liquid refrigerant is pumped to the generator and the rest goes to the evaporator, reaching an evaporating pressure by the expansion device. The process inside the cycle can be shown in figure 2.

The necessary heat input to the generator (Qg) is

(2)

(3)

Q = m (h — h . )

g g g, out g, rn

The cooling capacity at the evaporator (Qe) is,

Q = m (h — h. )

Q0 „

Q

2

Figure 2. Ejector refrigeration subsystem

e e e, out e, in

At the ejector, the energy balance at the mixing point is written as,

(4)

(mg + me) ■ hm = me ■ he + mg ■ hg, exp

COP

ejc

(5)

Q

Q,

The efficiency of the ejector system can be expressed by both an entrainment ratio (•, a ratio between the evaporation mass flow rate and the generation mass flow rate) and a coefficient of performance (COPejc). Neglecting the electricity supply to the pump, the COP of the ejector refrigeration system is defined as the ratio between cooling capacity and necessary heat input.

The ejector is the key component of the refrigeration subsystem, it is used to maintain the pressure difference between the condenser and the evaporator; the better the ejector, the higher the system performance. From the mass conservation, the impulse law and an energy balance around the ejector, the entrainment ratio (*) can also be written as:

# = me 1 mg = (cg 1 cc)“1 = [(h3 — h4 )/(h6 — h5 F -1 (6)

The ejector efficiency selected for this paper is typical for ejector performance as reported in the literature of Lundqvist (1987).

System Performances

Energy inputs to the system are heat to the generator, electricity to the pumps (both in the refrigeration subsystem and in the solar collector subsystem) and electricity to the auxiliary heater. The performance of the system can be defined as the system thermal ratio (STR).

(7)

STR =

Qsu

the heat supply to the STR, the ideal system is simply written as the

The electricity input to the pump is very small comparing to generator, thus it is generally neglected when calculating the performance can be shown as the system thermal ratio, which product of the collector efficiency and the COPejc.

(8)

STRideai = n„ • COPjJc

BRDF results and validation

As detailed in Andersen (2002), three types of graphical representations were developed to provide various visualization possibilities of the transmitted or reflected light distribution features, in addition to a recombined view of the six calibrated images, gathering the latter into a unique orthogonal projection:

• the projection of the BT(R)DF values on a virtual hemisphere, allowing a precise anal­ysis of the angular distribution;

• a photometric solid, representing the BT(R)DF data in spherical coordinates with grow­ing radii and lighter colors for higher values, illustrated in Figure 6;

• several section views of this solid, providing an accurate display of the numerical values distribution.

BRDF visualization*: photometric solid (hemispherical light reflectance* = 0,5)

BRDF visualization*: photometric solid (hemispherical light reflectance* = 0.01)

.

Figure 6: BRDF representation as a photometric solid.

285

270 I 10.05 255

‘ I.

(a) Opalescentplexiglas, (вг, фг) = (40P, CP) (b) Holographic film (HOE), (вг, фг) = (OP, OP)

An in-depth validation of both BTDF and BRDF was conducted, based on different ap­proaches (Andersen, 2004): [13]

• bidirectional measurements of systems presenting a known symmetry and verification against standard luminance-meter data or analytical calculations;

• empirical validation based on bidirectional measurements comparisons between dif­ferent devices; in case of disagreement, however, no conclusion can be established;

• assessment of hemispherical optical properties by integrating BT(R)DF data over the whole hemisphere and comparison to Ulbricht sphere measurements (Commission Internationale de l’Eclairage, 1998);

• comparison of monitored data with ray-tracing simulations to achieve a higher level of details in the BT(R)DF behaviour assessment.

These studies led to a relative error on BT(R)DF data of only 10%, allowing to confirm the high accuracy and reliability of this novel device.

Conclusions

This paper presents the conception and construction of an innovative, time-efficient bidi­rectional goniophotometer based on digital imaging techniques and combining BTDF and BRDF assessments. To allow reflection measurements, a controlled passage of the incident beam into the measurement space was created, minimizing parasitic reflections around the sample. Openings in the detection screen for the situations where it obstructs the incom­ing light flux were also required, made as small as possible to restrict the produced blind zones; to remove these elliptic covers, a motorized extraction and repositioning system was developed and tested successfully

This design proved efficient and reliable, for both the light beam penetration into the mea­surement space and the passage through the obstructing screen. The high accuracy achieved for BTDF assessments was checked to be kept for BRDF measurements as well, placing re­liance on the assumptions made in the construction of the instrument.

Acknowledgements

This work was supported by the Swiss Federal Institute of Technology (EPFL) and the Com­mission for Technology and Innovation (CTI). The authors wish to thank Pierre Loesch and Serge Bringolf for their contribution in the photogoniometer’s mechanical development.

Monthly and daily performance

3000

2500

2000 :

500

0

Figure 6 shows the daily delivery water temperature (OTL), ambient tem­perature, and tank bounding walls tem­peratures. It is ob­served that there is a fall of water tem­perature having co­incidence with the occurence of water consumption (curve ). This fall

is in the order of 2 Figure 6: Daily performance of the facade for domestic hot water pro­to 6 depending duction. Day March 11th. Barcelona climate. Consumption profile 1

on the water flow

required (at 19 hours, the larger consumption produces the larger drop of temperature). Numerical monthly results obtained for both climatic conditions, are shown in Table 6. represents the mean monthly outlet water temperature from tank (delivery temperature).

Barcelona:

Month QLOADMJ/ni~}

OTL°C

Sol,

ih

Geneve:

QLOADMJ/in-}

OTL°C]

Sol,

ih

1

158.01

26.07

0.38

0.46

72.53

20.14

0.17

0.37

2

145.73

26.15

0.38

0.46

76.38

20.83

0.2

0.38

3

166.73

26.49

0.4

0.47

134.33

24.21

0.32

0.43

4

158.7

26.29

0.39

0.49

129.71

24.16

0.32

0.46

5

161.22

25.93

0.38

0.5

140.07

24.43

0.33

0.49

6

162.73

26.53

0.4

0.53

160.31

26.24

0.39

0.52

7

175.19

26.94

0.41

0.55

183.16

27.57

0.44

0.53

8

189.16

28.25

0.45

0.55

181.99

27.43

0.43

0.52

9

191.02

28.82

0.47

0.53

172.76

27.36

0.43

0.5

10

196.6

28.65

0.47

0.5

141.12

24.79

0.34

0.47

11

175.33

27.63

0.43

0.48

72.2

20.23

0.18

0.39

12

156.18

26.51

0.38

0.45

61.23

19.55

0.15

0.34

Total

2036.59

27.02

0.41

0.5

1525.79

23.91

0.31

0.45

Table 6: Monthly perfomance for both climatic conditions. Consumption profile 1, stratifica­tion considered represented by 5 nodes_________________________________________

How those issues were addressed —

The architectural approach was an holistic one, from environmentally responsible design, to a healthy building for its users, with positive commitments at all levels. The building not only addresses ecological and energy issues, but health, economy and community.

From a global perspective, the intent was to whatever possible, within the given constraints of an existing building and limited budget, to contribute to the reduction of greenhouse gases. This has been achieved by an array of ESD measures, including solar passive design where possible, supplemented with energy efficient and energy reducing means for active systems where able.

From a national perspective, there was a conscious effort to utilise Australian products in preference to imported products where available, to be a leader in local government, and to contribute to society in a socially responsible way.

From a local perspective, local and regional tradespersons, suppliers and manufacturers were given preference where possible, including the local Builder who won the tender over several larger metropolitan based firms.

The most significant of these is the Geothermal HVAC system, utilising the stability of the below-ground temperature. In a climate with winter frosts and sub-zero temperatures, over 40 degree heat in summer, and a diurnal range in spring and autumn of 20 degrees plus, the geothermal, despite its initial capital cost, incurred a payback of just over four years.

Timescale

The project was engaged in November, 1998, and the first sketch plans presented to Council prior to Christmas that year. Construction documentation was completed for the Tendering process by April, 1999, with construction commencing in May, 1999.

Work was completed in February, 2000 with an official opening by the Premier of NSW soon after. Awards include Banksia 2001 Environmental Awards nomination, the Inaugural 2001 Green Building Awards Bronze Medal, and 2001 RAIA Country Division Awards.

Reference plant and parabolic trough model

In order to examine the feasibility of integrating external solar heat into the water steam cycle of a power plant, a model based on an existing reference plant was de­fined. This reference plant is a typical fossil fueled conventional steam power plant with an output of 393 MW and 7 feed-water preheating stages. Using APROS a com­prehensive model of this power plant was built-up, adapted and parameterized. Es­pecially the parameterizing is a very time-consuming step due the great amount of lay-out data like geometries, isometries, geodetic elevations, thermodynamic data, valve characteristics and automation concepts that have to be integrated into the model.

The power plant has been subdivided into more than 500 single components. For each component the one dimensional unsteady differential equations for the conser­vation of mass, momentum and energy are solved. Heat transfer, heat capacity of solid walls and two phase flow phenomena are taken into account. On the basis of this model several different kinds of plant configurations with or without an external heat source have been simulated. Careful calibration of the model has been carried out to meet the steady-state design and guarantee values.

To model the parabolic trough collector in APROS, design values calculated by steady state simulations combined with data taken from the literature were used [1,9]. The generation of steam with a heat capacity of 60 MW at noon in July was set as boundary condition for the collector design. The simulation is done with parabolic collectors of the type LS-3 (see figure 2) with a length of 100 m. Ten of them are added up to a 1000 m collector line. According to this, 15 lines in parallel are neces­sary to provide a peak load of 60 MW required for the simulation.

The collector efficiency is assumed to be constant at 67 %. This simplification is of sufficient accurancy for the simulations carried out in this paper, but will be corrected in future work. The collector feedwater pump control guarantees that the water pumped through the absorber tubes gets vaporized and superheated to a constant temperature of 380°C.

Cooling system configuration and modelling

• System configuration:

Figure 2 shows the configuration of the three heat reservoir cooling system. The main section of the solar cooling system can be divided into three flow loops: The solar collector flow loop, the ejector and cooling flow loop. The three flow loops are crossed by the same working fluid (R134a, R123). The choice of the refrigerants was subordinate to the saturation curve shape and to a certain extent to the knowledge of their thermodynamic and thermophysical properties.

The superheated vapour produced in the solar collector is sent to the ejector flow loop where the driving effect is produced. In our case, the intensity of the solar radiation has a direct influence on the vapour mass throughput of fluid at the collector exit and implies an intermittent flow (e. g. the solar collector in Tamanrasset (far south) will record a value of mass flow rate higher than in Algiers (north)). At the condenser exit the fluid flow separates into the driving massflowrate and the secondary massflowrate related each other by the entrainment ratio ra.

In Figure 3, the double Rankine cycle is illustrated in a logP-h thermodynamic diagram. It’s a combination of two basic Rankine cycles. The R123 and the R134a refrigerants have a positive — slope saturated vapor line. So we do not need important superheating compared to negative slope saturated line fluids where the isentropic expansion 1-2’ can induce vapor condensation that could affect the ejector performances.

Fig.1 the solar cooling ejector system

RESULTS OF THE DATA MONITORING

The DEC — plant is continuously monitored by registration of about 70 measuring points. All the important process information like temperature, humidity, volume-flow, operation signals of the humidifiers, ventilators, desiccant and heat recovery wheel, volume-flow controller and the position of the air duct flaps are registered in a 10 s — time step. For the evaluation one minute average values are used.

The monitoring campaign gives the chance to have a detailed look on non-ideal plant behaviour and offers thereby the basis for optimisation. Furthermore the monitoring suits to the purpose of evaluating the customer satisfaction and of evaluating the energy performance. The important value for the customer satisfaction are the room conditions. The energy performance is characterised by electricity and fuel consumption, solar gains and the efficiency of the heating and cooling performance.

Figure 3 shows a sketch of the flow sheet of the desiccant cooling plant including the ducts to the room.

1.0 Accuracy of measurements

Evaluating the monitored data some systematic errors came up:

• volume-flow measurement of small volume-flows: The volume-flow-meter has to work reliably in a range of 500 to 10200 mF/h. This is almost impossible for one single volume-flow-meter. Therefore the small volume-flows values have a large error margin.

• Humidity and temperature measurement in air ducts with rotating elements and little turbulence gives only the temperature information of the measuring position. For the introduced SDEC plant the dehumidification of the desiccant wheel can be analysed only in tendencies. The absolute values of adsorbed water vapour are not representative because of the strong influence of the rotation on local temperature and humidity distribution /6/.

• As inlet temperature the ambient temperature, measured at the north side of the building, is used. Meanwhile additional temperature measurements showed that the inlet temperature at the entrance of the plant is higher than the north side temperature. The reason for this temperature difference are natural convection effects at the east facade of the building, where the inlet duct for the desiccant plant is located.

1.1 Plant operation

The yearly plant operation hours depend on the using times of the rooms. Both rooms are used as meeting rooms and therefore they are discontinuously occupied. In 2002 the total amount of operation hours amounts up to 1335 hours, in 2003 up to 1289 hours.

Table 1 shows the operation hours and real using hours of both rooms. Operation hours means the hours the plant is running and air conditioning the room. Using hours or hours of occupancy means the hours where the rooms are really used by people for meetings. In 2002 for example the "Cafeteria” was air conditioned for 681 hours, but used "only” 506 hours. This difference is caused by the start — up period of the plant, where the rooms have to reach the comfort conditions.

operation hours

Cafeteria on

Cafeteria

occupied

Sitzungssaal on

Sitzungssaal

occupied

01 — 12/2002

1335

681

506

901

623

01 — 12/2003

1289

517

336

962

731

table 1: comparison of operation hours and hours of occupancy, 2002/ 2003

For evaluating the difference the facility manager of the IHK SO is keeping note of the hours of occupancy.

1.2 Room conditions and user satisfaction

The room conditions can be compared against the requirements of the German standard DIN 1946 part II /13/. This standard defines temperature and humidity thresholds for indoor comfort. These requirements, given by the red frame, can be visualised in a temperature — humidity — diagram, as shown exemplary in figure 3.

figure 3: comfort area (cp. DIN 1946 part II) of the “Sitzungssaal”, 1 — 12/ 2003

The points represent 1-minute average values of return air temperature and humidity in the rooms. The return air values represent the whole room situation. Therefore they are controlled and evaluated. The different seasons of the year are illustrated by different symbols. Most of the measured return air values are inside the comfort area. For the summer and spring quarter some data points are lying outside. These data points are mainly exceeding the limit of 11,5 g/ kg absolute humidity but keeping the limit of 65 % relative humidity. This situation is caused by the control design, which was controlling only a threshold of relative humidity. This control design was changed in July 2003. Now there is included an absolute threshold limit.

Characterising the comfort conditions the hours of threshold exceedance were evaluated. Table 2 shows the using hours and exceeding hours of both rooms for 2002 and 2003 separately. The calculated ratio shows clearly the more difficult situation of the "Cafeteria”. Their glazed facade is orientated to east, south and west and therefore extremely influenced by external loads caused by irradiation.

Cafeteria

Occupied

th]

limits

exceeded

th]

ratio

Sitzungssaal

occupied

th]

limits

exceeded

th]

ratio

01 — 12/2002

506

150

0.30

623

56

0.09

01 — 12/2003

336

87

0.26

731

95

0.13

table 2: absolute and relative hours of limit exceedance, 01/2002 — 12/ 2003

The main part of the exceeding hours are caused by humidity limit exceedance. As mentioned above there where some reasons within the control design. Analysing the plant performance, some other reasons could be found out. For example, the assumed efficiency of the heat recovery wheel was not reached. Looking for reasons, it was realised that the two rotating wheels, i. e. dehumidifer wheel and heat recovery wheel, have to rotate in opposite direction in the cooling case /6/ and in same direction in the winter case. Caused by the german climate the wheels were optimised for the winter case. The influence of changing the direction of rotation will be analysed in further investigations in 2004.

0 10 20 30 40 60 60 70 80 80 100

of operation hours

figure 4: cumulative frequency curves for ambient and room temperature 2002 and 2003

Figure 4 illustrates the comparison of the years 2002 and 2003 in an other way. The cumulative frequency curves are evaluated using 5 — minute-average temperature values only for the operation times of the plant. As room temperature the exhaust air temperature of the air conditioned room is used. For the times, where both rooms are used the temperature of mixed air streams is chosen. Figure 4 illustrate the percentage of temperature ranges within the operation time.

Looking at the ambient temperature of 2002 and 2003 it can be seen, that in 2003 at 25 % of the operation time the ambient air temperature was higher than 25 °C, whereas in 2002 this was the case for 11 % of the time only. Looking at the room temperature, in 2002 2.5 % of the operation time the limit of 27 °C was exceeded. In 2003 this was in 5 % of the operation time the case.

SHAPE * MERGEFORMAT

In /14/ it was evaluated that in Freiburg the average ambient air temperature during summer1 2002 was 1.5 K higher and during summer1 2003 5.2 K higher than in the Test Reference Year, which was used for simulation. Comparing the results with the promised 0 — 2 % operation time exceeding the temperature limit of 27 °C of the simulation study, the realised conditions are very satisfying.

1.3 electricity consumption

The total electricity consumption for the whole monitoring period was about 26649 kWh. This value is measured in total for all plant consumers by a central electric power meter.

07 — 12/2001

01 — 12/ 2002

01 — 12/ 2003

Total

electricity consumption

6652

9577

10420

26649

table 3: electricity consumption 07/ 2001 — 12/ 2003

The electricity consumption is in general far higher than estimated. Three main reasons

were found to be responsible for the high consumption /15/.

• High stand-by electricity consumption, which varies between 0,2 and 0,5 kW. The higher value is caused by security functions which are active in winter. For the whole year this means a standby electricity consumption of approx. 2600 kWh.

• Fan efficiency was overestimated. The calculations done in the phase of plant design act on the assumption of a constant overall fan efficiency of 0.6 (hydraulic/electric). This assumption is right for high volume flows; in this case the efficiency can be even higher. But the volume flow is varying according to the room demand and is often quite low. In 2002 and 2003 the inlet volume flow is at more than 60 % of the operation time lower than 4000 m3/h. At 4000 m3/h the fan efficiency is about 0.45 and at 2000 m3/h for example only 0.25.

inlet volume now in m’/h/10

figure 5: el. consumption before and after changing air duct pressure control (minutely values)

Constant air duct pressure control caused at low volume flows higher air duct pressure than necessary for delivering the air. The higher pressure must be generated by increasing the frequency and therefore the electricity consumption. This control was changed in summer 2003. Now the demanded air duct pressure is implemented as a function of the volume flow. Figure 5 shows the results of the changed control. The electricity consumption per inlet volume flow decreased significantly.

The reasons found for the high electricity consumption are not caused by solar components of the system. This means that high stand-by consumption and constant air duct pressure control could in the same way be also a problem for conventional ventilation systems. For both the solar desiccant system as well as the conventional air handling unit there is a potential for optimisation in general.

Discussing the energetic performance one should compare the solar DEC plant with a reference system. The reference system must be a ventilation system serving the same rooms with the same comfort, therefore a compression chiller is needed. Calculating the reference system return air humidification and heat recovery is used to minimise the refrigerating capacity of the chiller. The COP of the chiller is assumed to be 4. For the reference system a nominal pressure drop of 727 Pa and for the solar DEC system a value of 1382 Pa was assumed. The higher nominal pressure drop of the solar DEC system is caused by the desiccant wheel, the second humidifier and the collector field. The duct pressure drops are the same.

figure 6: comparison of electricity consumption of a solar DEC plant and a compression chiller driven reference system

For calculating the electricity consumption the above mentioned findings concerning like stand-by consumption, volume flow dependent efficiency and air duct pressure control were considered in calculation. Therefore, in figure 6, the calculated electricity consumption of the solar DEC system is similar to measured values (cooling mode, 2002).

Comparing only the electricity consumption of fans and pumps, as ca be seen in figure 6, the reference system needs less electricity. This is caused by lower pressure drops and less components. Calculating the electricity consumption for the reference system with compression chiller (COP = 4), the ambient and room conditions given in figure 1 were assumed. Part of the refrigerating capacity is covered by enthalpy recovery. The necessary heating capacity was neglected for the discussion.

As figure 6 clearly shows the electricity consumption of a reference system with compression chiller at the assumed conditions would be definitively higher. Of course it has to be taken into consideration that caused by the not optimised direction of rotation during the summer season in the last years, the assumed inlet air conditions could not be always realised with the solar DEC system.

1.4 Collector performance

The yearly air collector performance is characterised by total specific irradiation and the specific collector gains. For the evaluation of a system without any storage and irregular operation, the irradiation in times of operation is also important to know.

Therefore the efficiency, ■p, for this solar autonomous system is defined in two different ways.

П, оы = (1)

QIRR, total

n. = —Qolaa:— (2)

operation

IRR, operation

Equation (1) defines the efficiency by dividing the gained solar energy by the total irradiation. Equation (2) defines the efficiency by dividing the gained solar energy by the amount of irradiation within the times, where the plant was operating.

In table 4 these values are given for 2002 and 2003. The value of the total collector efficiency is quite low, because of the low amount of yearly plant running hours. The collector efficiency in operation times is ranging between 20 and 25 %. Taking into account, that the collector gains are only usable in the desiccant and the heating operation mode, this is an acceptable result.

total irradiation [kWh/ m2]

Irradiation plant on [kWh/ m2]

total collector gain

[kWh/ m2]

total collector efficiency

collector efficiency plant on

01 — 12/ 2002

1092

371

74

0.07

0.20

01 — 12/2003

1296

414

100.8

0.08

0.24

table 4: collector performance values

є

collector

32,4

(3)

solar ,2001—2003

P

EL, coll

Interesting for primary energy aspects is the question of used ventilator electricity for "collecting” the solar gains. Therefore an electricity coefficient is defined by dividing the solar gains by the collector fan electricity consumption (equation 3). This was done for a period between July 2001 and end of April 2003. The electricity coefficient is about 32.4.

It is important to mention that the electricity consumption of the collector fan is also influenced by the pressure drop of the desiccant wheel. Its pressure drop is at least in the same range as the pressure drop of the whole collector array. Therefore it can be estimated that the electricity factor including primary energy consideration would be at least double as high if only the electricity consumption due to the solar collector pressure drop is considered.

In 2002 and 2003 nearly 19 % of the collector gains were used for heating purposes. In the heating case between 15 and 18 % of the necessary energy input could be covered by this solar input.

1.5 Cooling performance

The cooling energy is calculated by balancing the input air stream between ambient and input conditions. The driving energy input for the cooling case is solely covered by the solar air collector gains. Therefore the solar fraction for the cooling case is 100%.

The cooling process can be characterised by the "coefficient of performance”, COPthermai. This value points out the ratio of useful and invested energy.

COP _ Qcoolmg Vinlet P! hambient hinlet ) (4)

thermal (4)

solar, cooling solar, cooling

Table 5 shows the values of the cooling performance for the whole monitoring period. The thermal COP varies between 21 and 38 %. The difference between the average thermal COP value of 2002 and 2003 can be explained by different ambient conditions (cp. Figure 4) and by changed volume flow distribution of the inlet volume flow.

total cooling energy plant [kWh]

collector gains used for cooling [kWh]

COP, thermal

01 — 12/ 2002

1263

6039

0.21

01 — 10/ 2003

3068

7996

0.38

table 5: cooling performance values

The thermal COP increases with increasing ambient temperature. This is caused by the greater temperature difference between ambient and room. Looking at the COP as function of the inlet volume flow, it was realised, that the COP increases with increasing volume flow. In 2003 the average inlet volume flow was higher than in 2002. This effect is caused by the specific configuration of the solar DEC plant. There are times where only a small inlet volume flow is needed. In case of high irradiation the energy given by the collector field is more than needed for regeneration.

A COP comparing the cooling capacity with the electricity consumption therefore required is not yet available on an annual basis. But looking at the assumed conditions of figure 1 and the results of figure 6 for a volume flow of 8000 m3/h the electricity related COP for the reference system would be about 1.5 and the one of the solar DEC system 5.2.

As already mentioned, in the summer case the heat recovery wheel and the desiccant wheel should rotate in opposite directions /6/. This will be realised within the cooling season 2004. A significant increase of the heat recovery efficiency is expected. The results will be presented in future publications.

Method

1.1 Experimental Set-up

Measurements were executed in the Visual Comfort Evaluation (VCE) set-up which consists of a 1:5 scale model of a 3.6 x 5.4 x 3 m3 office room with an artificial overcast sky, see [1]. The inner walls are matt black (RAL 9005) to exclude the influence of internal reflections on the effect of the contrast region. A hole in the long side of 0.3 x 0.6 m2 simulates the window. The wall next to the window is painted matt white (RAL 9010). All materials are positioned to the right side of the window, and if the materials were not high enough they were elevated using a piece of string, see figure 1.a.

All luminance measurements were performed with a Minolta LS-110. The top-angle is 1/3 degree. At a minimum distance of 1014 mm, the diameter of the measured spot is 4.8 mm. The measurements were taken at a height halfway between the top and the bottom of the window. On the right-hand side of the wall measurements were taken every 2.5 cm in the horizontal direction. Two measurements were taken in the transition region, and one in the window along the line given in figure 1.b.

Figure 1: Experimental set-up (a) and measurement points given in the vertical cross section (b)