Category Archives: EuroSun2008-7

Organisation of the Project

The project is based on the skills of each partner working within four pools.

• Pool 1 provide a general expertise and a thermodynamic analysis of the processes

• Pool 2 must study the desiccant processes

• Pool 3 is devoted to absorption process

• Pool 4 focuses on thermo-chemical processes

The pool 1 is in charge of the comparison and the analysis of the three production processes studied by the pools 2, 3 and 4. Poles 2, 3 and 4 will aim to produce a various tools based on the results of the analysis of the pole on 1 and their own results.

2. Experimental facilities

A fundamental stage before the use of the simulation tools for new projects is their validation. This complex procedure includes many steps as an inter software comparison, an experimental validation to compare simulation results with experimental measurements and a parametric sensitivity analysis. The data generated by the experimental prototypes are an important asset to validate our tools. Two procedures are planned. We will first use experimental prototypes for the validation of the components models, of the systems and of the coupling between the system and the building and then we will study real installations for the validation of coupling between the system, the building and the users.

The first experimental results

We decided to present the 2nd of June 2008; it was a very sunny day. We can see the evolutions of temperatures and powers on fig. 2 and 3.

At 8 AM, the outlet temperature of the solar collectors increases gradually to exceed the temperature of the hot water tank around 9 AM. At the same moment the three ways valve on the solar loop between solar collectors and the hot tank opens to warm the water of the tank. At 10 AM, the students come in the first classroom, the air temperature inside the classroom increases slightly to reach its maximum, 24°C at about 11:30 AM.

Around 11 AM, the hot tank temperature reaches 80°C, in the same time the absorption chillers starts.

At the beginning, the powers of the generator and the cooling tower are very high as we can see on the fig. 3. These powers dwindle quickly during the first hour and even more during all the functioning time. We can notice that the chiller begins to produce cold water about 10 minutes after the starting up and this refrigerated power increases slowly the next 10 minutes. Around 30 minutes later of functioning the refrigerated power stabilizes in 23 kW. At the same time the thermal COP follows closely this evolution particularly at the beginning.

We notice that the electric COP remains around 3.8 throughout the functioning of the chiller.

After less than one hour of functioning, the water of the cold tank is distributed in the first classroom. Between 12 PM and 1 PM, the students leave the classroom, so the air temperature inside the classroom decreases quickly as we can see on the fig. 2. At 1 PM, the students go back

image557 image558

to the classroom, so the air temperature remains around 22°C for an outside temperature of 28°C. The difference between the outside and the inside air temperature is 6°C so the comfort objective is reached.

Fig. 2 et 3: Evolution of the temperatures and powers in each components

At 4:15 PM, the temperature of the water inside the hot tank dwindles to 65°C so the generator pump stops. The generator of the chiller is not supplied by hot water but thanks to its inertia of the machine, the chiller goes on to produce cold water. This fact explains the peak of the electric and thermal COP at the end of the daily operation of the machine. Ten minutes after stopping the pump generator, the chiller stops. Finally, after 1 hour shutdown, the distribution of cold into the classroom stops because the cold tank temperature falls to 20°C.

Variation of collector field area and storage size

The system was simulated with different numbers of collectors and varying sizes ice storage. The parameters variation is shown in table 2.

Table 2 Simulated collector’s area [m2], storage volume [m3] and capacity [kWh]

Number of collectors

11

13

15

18

20

Net aperture area [m2]

35,75

42,25

48,75

58,50

65,00

Size of storage vessel [m3]

1,5

2,0

2,5

3,0

Latent capacity [kWh]

70,4

93,9

117,4

140,8

3,0m3

2,5m3

2,0m3

1,5m3

4 V

4

4 — x

* m

23

21

19

17

15

Подпись: Number of CollectorsПодпись: CO 4—' CD image61210 12 14 16 18 20 22

Number of Collectors

Fig. 9 Solar Fraction of the system with differed Fig. до Overall efficiency of the system with different

storage volume storage volume

Following the results of the simulation (timestep = 1 min) are presented in the form of diagrams.

The simulations highlighted that the system’s solar fraction, for the given process rises with the collector field size, ranging from approximately 21 to 28% depending on the storage size. At the same time the overall efficiency of the solar system drops from about 22 to 16%. The latter is mainly due to defocusing of the collector (dumping energy) when the solar heat input exceeds the

needs (see Fig. 9). Moreover the influence of the storage size shows a milder effect on the solar fraction and system efficiency than the size of the collector field. Nevertheless it is to mention, that the influence of the storage size increases with larger collector field’s areas due to the fact that part of the additional heat delivered by the bigger collector field can be stored on the “cold side” of the system. With smaller collector fields a large storage is not needed, since in most cases the smaller amount of produced cold will be extracted from the solar cooling system before the storage will be completely charged (avoiding collector defocus). Furthermore the analysis highlighted that the

Подпись: Fig. 11 Ratio of dumped energy and theoretical energy gain of the collector maximum SF achievable (about 28%) for the system concept and application is slightly smaller than for conventional air-conditioning solar assisted systems.

Solar thermal plant

Three days with different conditions have been simulated and compared with measurements. Table 2 summarizes the results by comparing the energy measured and simulated for the different days:

Table 2: Summary of the measured and simulated results of the solar thermal plant

Qsol

Qcol_m

Qcol sim

Qload m

Qload_sim

Dev. col

Dev. load

kWh/m2

kWh/m2

kWh/m2

kWh/m2

kWh/m2

%

%

10th May

8.38

4.69

4.55

4.04

4.47

3.1

10.8

5th July

8.04

4.32

4.17

3.92

4.04

3.2

3.1

11Th July

4.93

2.57

2.27

2.10

2.08

11.4

0.8

2.1.1. Collector loop

The results show a good agreement between measured and simulated values with an acceptable deviation. Nevertheless, it should be mentioned that when the irradiation is fluctuating, the results are not so good (11.4% deviation in comparison to measured values). This might be explained by

the fact that the values used to perform the simulation are mean values over 5 minutes-period. The values of volume flow rate and irradiation are therefore not “real”, that is why one can observe these peaks in the collector outputs when the volume flow rate is lower. These problems can easily be avoided by using instantaneous values and by reducing the time step.

Indirect Evaporative Cooler Effectiveness

In the previous paper, White et. al. [1] used the default effectiveness factors provided in the TRNSYS type 757 indirect evaporative cooler model, being in the range 0.3 to 0.5 depending on the primary air dry bulb temperature and the secondary air wet bulb temperature. However,

ASHRAE [7] suggests that indirect heat exchangers can achieve a 60 to 80% approach to the secondary air wet-bulb temperature at a secondary to primary air flow ratio down to 0.6:1.0.

The potential benefit of increasing the effectiveness of the indirect evaporative cooler was first investigated in a series of stand-alone, steady-state simulations of (i) a two stage indirect/direct evaporative cooling process and (ii) a single stage direct evaporative cooling process. The performance of the two stage indirect/ direct process was evaluated by how much colder supply air could be made, compared with that from a single stage direct evaporative cooling process. The results of this analysis are presented in Figure 2 for a range of inlet air conditions.

Подпись: 9 і

Подпись:Подпись: О V Подпись: <u a. Є v Подпись:Подпись: 0.005Подпись: 0.01Подпись: 0.015Подпись: 0.02Подпись: 0.025Подпись: 5Подпись: 4Подпись: 0Подпись:image1418

Q.

7 0

0.03

At high primary air humidity, the inclusion of an indirect evaporative cooling step has limited impact on the achievable supply air temperature compared with a simple direct evaporative cooler, irrespective of the indirect evaporative cooler effectiveness. However, at low primary air humidity, the inclusion of an indirect evaporative cooling step significantly reduces the supply air temperature compared with a direct evaporative cooler, and increasing the indirect evaporative cooler effectiveness from 50% to 70% almost doubles the impact of the indirect evaporative cooling step.

Given that the heart of the desiccant cooling process is the reduction of primary air humidity, the use of a two stage indirect/ direct evaporative cooling process appears to be sensible. The additional temperature reduction can potentially expand the number of hours over which backup cooling is not required, and increase the thermal cooling from the same quantity of dehumidified air (with resulting increase in coefficient of performance (COP)).

The impact of increasing the effectiveness of the indirect evaporative cooler was further investigated in TRNSYS simulations of the desiccant cooled building described in Section 2, but with a fixed desiccant regeneration temperature of 60°C. The resulting predictions of thermal conditions in the occupied space, for each of the two assumed values of indirect evaporative cooler effectiveness factor, are illustrated in Figure 3.

image142

Figure 3 suggests that, under the stated assumptions, a target zone temperature of 26°C is exceeded for less than 20 hours, with air change rates of 3.71hr-1 and 4.53hr-1 for 70% effectiveness and default TRNSYS model 757 effectiveness values respectively.

The increased indirect heat exchanger effectiveness, and consequent decrease in required air flow, leads to reduced fan parasitic energy consumption and reduced equipment size and cost. The reduced flow also reduces the amount of regeneration air to be heated with an expected improvement in the coefficient of performance.

Facility elements model

For the heat production there is a solar camp with high efficiency collectors, with the following parameters for their performance curve:

Maximum efficiency: k0 = 0,8

Loss Coefficient: ki = 4 W/m2K = 14,4 kJ/hm2K

They have been chosen realist values for the flat plate collectors performance. Note there are commercial options with better performance, but the aim of the project is to evaluate an installation as much conventional as possible. It is supposed the panels are perfectly oriented to the south with a slope related to the horizontal of 45°. Note that for Valladolid the latitude is 42°. They are placed in such way that they are free from shadows.

To estimate the area, it was supposed an habitual design ratio for solar absorption installations, of 4 squared metros for each kilowatt of power on the evaporator (SACE). And so, to have a facility with 7 kW, is necessary to have collectors area of 28 m2 to operate properly the installation. The flow for the generator camp selected was 50 l/hm2, that in an habitual value for design and test, what made the solar pumps flow to be 1.400 l/h. The storage has been modelled using a stratified tank. It has been used the bigger amount of nodes allowed by the model (15), with a lost coefficient of 0,8 W/m2K = 3 kJ/h m2K. Note that the performance of the installation depends very much on the stratification, and so if we take some models for the tank without stratification, the lack of variation on the temperature profile makes the installation performance to be very inferior.

An habitual value dimensioning is 50 l/m2 what gives an autonomy of 1 hour. Note that an overdimensioning of the storage capacity, overall on real facilities, can lead to need to much time to achieve the required temperature setpoint, and on the contrary to have a low value leads to a reduced autonomy. Other thing to take into account is that the delay between the inlet temperature respect to the outlet one is 2 hour. The absorption chiller has been modelled by means of the type 107 with the working curves corresponding to a WFC-10 [4] adapted to a power of de 7 kW what is the maximum power demanded.

The calculation period considered covers from 1st June until 31st September.

Monthly performances

By means of the collected data, mean performances figures and energy savings system were calculated also with the aim to compare such system with a reference conventional one. The energy performance indicators used for the evaluation are presented here below.

The SFDEC is the fraction covered by the desiccant cycle to the total cooling energy delivered by the AHU and is calculated by equation (1). The monthly solar heat management efficiency describes the quantity of incident irradiation what is usefully utilized in the system and calculated by eq. (2):

Подпись: _ QHC 2 heat Подпись: (2)QCC1 + Qcc 2

sFDEC _ 1 q (!)

QDEC

The thermal COP HC2 of the desiccant AHU indicates the ratio between the cooling energy produced by the desiccant cycle and the regeneration heat delivered by the solar heating coil HC2, whereas the thermal COP HC2+HC1 includes the global heat amount coming both from the solar and heat recovery coil.

image335

image336 image337
image338

(3)

 

image339

(4)

 

QH

 

QHC 2 + Qh

 

image340

In order to estimate the energy saving obtained in comparison to a reference AHU, the primary energy consumption of both systems have been calculated with the general formula (5) taking in account for the reference unit also the necessary heating energy delivered from a gas boiler to achieve the desired temperature for the supply air. The primary energy ratio PER has been calculated by the formula (6) where it was assumed the same value of QAHU both for the desiccant and reference AHU.

PE _ Qel + Qre~heating

 

image341

(5)

 

(6)

 

Vd

 

П fossil

 

image342

Since the desiccant wheel and the additional coils cause higher pressure losses than in a conventional AHU, different electricity consumption for ventilation have considered for the two systems. In particular, in the calculation of the primary energy consumption for the reference system, the electricity consumption of the reference AHU was calculated according to the mentioned “Monitoring procedure of solar heating and cooling systems” of the IEA Task 38. More in detail, knowing the measured electricity consumption of the desiccant AHU and assuming the same efficiency of the fan motors at the same flow rates, the one of the reference AHU can be derived introducing a correction factor related to the pressure drops AP and to the volumetric flow rates :

image343
image344

(7)

 

image345

image346

The calculated electricity consumption for ventilation of the reference AHU amounts to 47% of the one of the desiccant cooling AHU. In the calculation of the primary energy consumption related to the cooling energy it also was assumed the same chiller performances of the one used in the DEC system.

+ + + Ч ++++^

2146; 43%

+++++*

,+++++-
■■++++■
k++++

image347
image348

The next figures show the cooling energy QAHu delivered to the building for ventilation and sensible cooling and the distribution of the cold production in the AHU. It can be seen that, for the considered time period, the fraction of the cooling energy covered by the desiccant cycle (QDEC) is averagely 42%. It is worth to note that, since the radiant ceiling power was not sufficient to meet the sensible load of the room, a lower supply air temperature (18-20 °C) was necessary to guarantee the indoor set point conditions. Also for this reason the second auxiliary cooling coil gives a considerable contribution (QBF2= 47%) to the cooling energy balance.

Подпись: Q BF1Q radiant ceiling ■ Q vent sens Я Q vent lat

Fig. 6. Distribution of the cooling energy delivered to the building (left) and produced in the AHU (right)

+++*

++++

+++++

+++++*

1363; 41%

k++++

image350 image351

In Figure 7 shows that the energy delivered from the sensible heat exchanger and the solar heating coil HC2 are similar, whereas the contribution of the condensation heat recovery coil HC1 is 22% to the total corresponding to the half of the energy delivered by the solar heating coil. In the same figure, also the distribution of electricity consumption for the whole system is reported.

Подпись: Q HC 2Q HX ret

Fig. 7. Distribution of regeneration heat for the desiccant cycle (left) and electricity consumptions (right)

The primary energy ratio for the desiccant unit, PER AHU is 1.03 kWhcold/kWhPE, whereas the one for the reference system amounts to 0.54 kWhcold/kWhPE. In other terms, the primary energy saving of the desiccant system is 47% for the considered time period.

In the following table some mean measured values and performance figures for this time period are summarized.

Table 2: Mean monthly performance figures — July 2008

1st International Congress on Heating, Cooling, and Buildings — 7th to 10th October, Lisbon — Portugal /

Supply ventilation air flow rate

1831

[kg/h]

Temperature of chilled water

12

[°C]

Return ventilation air flow rate

1746

[kg/h]

SFdec

42

[%]

Total cooling power to building

14.5

[kW]

Regeneration temperature

54

[°C]

Sensible load of building

9.6

[kW]

Regeneration flow rate ratio

2/3

[-]

Latent load of building

3.9

[kW]

Solar collector efficiency

40

[%]

Total cooling power of AHU

9.1

[kW]

П solar heat

26

[%]

Cooling energy to building

5028

[kWh]

COPth HC2

1.14

[-]

Cooling energy delivered by AHU

3374

[kWh]

COPth HC1+HC2

0.72

[-]

Cooling power of chiller

10.6

[kW]

PER ahu

1.03

[-]

Electrical COP of chiller

2.73

[-]

Primary Energy saving

47

[%]

3. Conclusions

The whole system is working fairly well in summer operation. First monitoring data have shown promising results. Some improvements of the solar system will be necessary in order to increase the solar heat management efficiency and consequently the dehumidification potential of the desiccant cycle. The contribution of the condensation heat coil to the energy required for the regeneration has given interesting results and will be further investigated. Some mechanical adjustments in the heat exchanger should minimize the problem of air infiltration from the return side to the supply side. Furthermore, the accuracy of the humidity measurements must be improved. Nevertheless, considerable primary energy savings have been reached.

Nomenclature

CC or BF

Cooling Coil

Qdec, ahu

Desiccant cooling Energy of AHU

HC

Heating Coil

SFdec

Cooling Energy Fraction covered by DEC cycle

DW

Desiccant Wheel

n Solar, heat

Solar heat management efficiency

HX

Sensible Heat Exchanger

COP th, HC

Thermal COP of DEC cycle

HU

Humidifier

PER ahu

Primary Energy Ratio of cooling energy from AHU

PE

Primary Energy

£ el

Conversion factor for electricity production = 0.4

^boiler

Boiler efficiency

£ fossil

Conversion factor for fossil fuel = 0.9

References

[1] Dhar P. L., Singh S. K., 2001. Studies on solid desiccant based hybrid air-conditioning systems — Applied Thermal Engineering 21, 119-134

[2] Mazzei P., Minichiello F., Palma D., 2005. HVAC dehumidification systems for thermal comfort: a critical review. Applied Thermal Engineering 25, 677-707

[3] Finocchiaro P., PhD Thesis “Analisi numerica di sistemi desiccant cooling alimentati ad energia solare per applicazioni in climi mediterranei” Dipartimento di Ricerche Energetiche ed Ambientali — Universita degli Studi di Palermo, 2007

[4] Beccali M., Finocchiaro P., Nocke B., Gioria S., “Solar desiccant cooling AHU coupled with chilled ceiling: description of a new installation at DREAM in Palermo”, Proc. of the 2° OTTI International Conference on Solar Air Conditioning, Tarragona (E), October 18th -19th., 2007, pp 389-394, ISBN 978-3-934681-61-3.

The AC-Sun, a new concept for air conditioning

Saren Minds1* and Klaus Ellehauge2

1 AC-Sun, Rudolfgaardsvej 19, DK-8260 Viby J, Denmark
2 Ellehauge & Kildemoes, Vestergade 48 H, 2s. tv., DK-8000 Aarhus C, Denmark
* Corresponding Author, smi@ac-sun. com

Abstract

This AC-Sun is in the process of developing a solar powered AC unit which only consumes 10% of the electricity of a conventional AC unit. The devise consists of a traditional thermal solar panel that heats water to temperatures between 75-95°C. An expander, based on a Rankine — process, uses the energy from the solar panels to operate a compressor, which through a Carnot — process cools air in traditional air coolers. The process uses water as cooling agents both in the expander and the compressor and therefore is environment friendly. The core of the research and development project is the patented process, which consists of leading the energy from the compressor outlet back to the steam temperature on the inlet duct of the expander. The internal heat recovery ideally improves the system efficiency for cooling purposes compared to other similar systems. The developing process comprises design and production of process equipment and processing forms for paddle vanes and turbine casing as well as internal heat exchangers. A world-wide patent for the principles behind the solution has been approved through a PCT testing. The proto type test is ongoing in — August / September 2008.

Keywords: AC-Sun, Solar Air Conditioning, Rankine process, expander, turbo compressor

1. History

The project AC-Sun started at year 2005 with the mission to develop and commercialize a thermal air conditioning solution driven by solar panels. The novelty of the product is great due to considerable power savings (a factor 10), free-of-pollution, silent cooling and better comfort.

The technical content consists of research and development of a turbo expander/ compressor which is working with water as a process agent under low pressure and temperatures, and where the heat production from low temperature thermal solar panels is transformed into a low temperature compressor, which is used for cooling of air in an air conditioning device. The success criterion is to exploit a series of thermal processes, which by them selves are well-known, but never previously have been combined to one functioning unit. The world-wide patent on the principles behind the solution is entering the national phase and the company has taken the process to a stage where proof-of-concept has been achieved.

Task E: Implementation

This last Task of the project is thought to spread the results obtained during all the three years, basing the dissemination on a group of reports, that explain best case examples for different kinds of the final use and objectives.

A number of guidelines with do’s and don’ts, experiences on testing and labelling will be published under the form of handbooks and downloadable documents from a web page that explains the work done and the results obtained.

In parallel some workshops and training courses are being organized to introduce possible users, engineers and designers in the use and possibilities of the technology. A first event of this kind was organized in Bolzano (Italy) the 10th of April, presenting the possibilities of district heating coupled TDHPs to more than twenty engineers and technicians.

3. Conclusions

After thirty six months of work, a complete overview on the TDHP market should be available and a useful tool for the implementation of this technology in the identified cases, which have proven to be economically and energetically feasible solutions compared with those conventionally used.

Different ways to distribute the information obtains have been thought:

• An internet website linked to the HPC-site, giving up to date information on

i. market overview with possible applications and examples of demonstration projects or cases;

ii. tools for sizing of thermally driven systems and calculating their energy and economic performances;

iii. good practice guide evaluating users’ experience in different projects. (www. annex34.org)

• A description of a standard to determine COPs of thermally driven heat pumps

• An online database about sorption materials for heat pumps and their properties

• A reference guide describing presently available

i. thermally driven systems with their applications;

ii. software tools, their application and users experience;

• A final report describing the work carried out under the Annex 34.

• At least two workshops with complete proceedings.

References

[1] Tozer, R., Gustafsson, M. et all. Dec2000. Absorption machines for heating and cooling in future energy systems — Final report HPP Annex24

[2] Web page of the institute for environment and sustainability. http://sunbird. jrc. it/pvgis/apps/radmonth. php? lang=en&map=europe

[3] Werner S. et al., December 2006. EcoHeatCool. The European heat market. http://www. euroheat. org/ecoheatcool/documents/Ecoheatcool WP1 Web. pdf

[4] Dalin P. et al., December 2006. EcoHeatCool. The European cold market. http://www. euroheat. org/ecoheatcool/documents/Ecoheatcool WP2 Web. pdf

[5] Hamring J. et al., June 2007. Review of Renewable Energy in global Energy Scenarios. IEA

[6] Rodriguez J. et al., June 2008. Bombas de Calor Activadas Termicamente. XIV Congreso Iberico y IX Congreso Iberoamericano de Energia Solar 2008

Public subsidies

The Plan for Renewable Energies in Spain assigned 348M€ for the period 2005 to 2010 to subsidize private investments in thermal solar energy installations. The corresponding budget is annually distributed by the local Government of each region.

Malaga belongs to the region of Andalusia, where last year’s subsidy reached up to 50% of the solar field’s installation cost, as shown in the graphics below. The possibility of obtaining a public subsidy for the installation was considered in the analysis, individually and combined with other economic saving to reflect its effect on the viability of the system.