Category Archives: EuroSun2008-10

. Previous Studies

Two previous studies have been undertaken investigating ISAHP systems. The first study involved developing a model in TRNSYS [12], a transient simulation program, to investigate the feasibility of the system, performing both a performance and cost analysis. The second study involved building a prototype of the system, and performing a range of constant temperature input tests, comparing the results with the simulated results. A brief description of each previous study is provided below.

1.1. Numerical Analysis

The first study was conducted in the Solar Calorimetry Lab by Freeman [3] and simulated the performance of indirect solar assisted heat pump using TRNSYS. The program used mostly component models developed with the TRNSYS software, but models were created for the heat pump, the natural convection heat exchanger, and the heat pump controller. A detailed description of the TRNSYS model, and a theoretical analysis and derivation of the steady-state vapour compression heat pump model is given by Freeman [3], and is briefly summarized in a previous paper by the authors [13].

Results of these previous studies predicted higher seasonal solar fractions than conventional Solar Domestic Hot Water (SDHW) systems. It was concluded that the ISAHP gathered more energy from the environment during marginal weather conditions, as well as during the winter when compared to either an SDHW or air-to-water heat pump systems. The study also found that the life cycle cost of the ISAHP system showed up to 29% savings over the SDHW system for major cities across Canada.

Dimensioning for the best cost benefit ratio

The conventional dimensioning of a solar heating system for one — or multi-family houses is commonly based on a trade-off between the solar fraction and the level of utilisation [2], whereas the customer and user might rather be interested in elements like additional cost and primary energy savings. To meet these requirements a different dimensioning directive was introduced [1][3]. The simulation program TRNSYS was coupled with the optimisation program GenOpt for the dimensioning in response to this directive. A short description and the resulting optimal dimensions of the MaxLean system concept are given in the following and serve as basis for the subsequent sensitivity analysis.

3.1 Dimensioning directive

The objective function minimised by the optimisation algorithm is the cost/benefit ratio described in Eq. (1):

a*I0 + BMaxLean Bref

a annuity factor

I0 total investment costs of the solar thermal system

BMaxLean annual operation costs of the MaxLean system concept (including the heating circuit)

Bref annual operation costs of the conventional reference system

Eprim, sav primary energy savings

As long as a solar heating system is not economically rewarding, the primary energy saving is the preferential merit on which value is laid. The additional cost for these primary energy savings consists of the additional investment cost minus the difference in operation costs between the conventional and the solar assisted heating system. The investment costs of a solar heating system are calculated by Eq. (2), (cf. [2]):

A ( V 6536

I0 = 2559€ + *368€ + 3983€I I Eq 2

m2 t m3 I 4′

Acoll flat-plate collector area

Vstor storage device capacity

The annual primary energy savings are calculated from the difference between the energy consumption of the solar and non-solar heating system, as well as the embodied energy of the solar system. (Thus, the embodied energy of the non-solar heating system —i. e. a water heater store— is not accounted for). The gas and electricity consumption is converted to primary energy consumptions. To obtain annual values of the embodied energy the total amount is divided by the service lifetime.

State of the art and development in Europe

Solar thermal plants in the district heating scale were pioneered in Sweden the beginning of the 80’s. Since then a significant number of plants have been built around Europe, both with diurnal and seasonal storage. The technology was developed to a mature stage during the 90’s when a multitude of plants were built around Europe. Most of the plants today are found in Austria, Denmark and Sweden. Also in Germany there are a number of plants in operation. The difference in land prices between the northern and central Europe is probably the reason, that in Sweden and Denmark the collector arrays have been mostly built as ground mounted, whereas in central Europe they are normally mounted on roofs. The large number of solar assisted block heating plants in Austria is mainly thanks to generous subsidies, which have been around 40% of investment costs. Another fact that has made it feasible there is that traditionally the block heating networks were shut down for the summer to avoid the high heat losses. A new summertime heat supply business could be started when investing in solar and could be seen as extra income. The situation is different from the northern European view, where summertime district heating (even though with bad thermal efficiency) is taken for granted, and solar thermal is seen only as a means to save fuels. [1-12]

Twin-Store-System with heat pump (system A)

Подпись: Fig. 1: Twin-Store-System with heat pump (system A)

In the Twin-Storage-System (see figure 1) a combistore (nominal volume 600 l) is used for hot water preparation (by means of a stainless steel tube) and for space heating. It is charged by the collector via an immersed heat exchanger located in the lower part. The auxiliary part of the combistore is charged by the heat pump. Another pressureless buffer store (nominal volume 800 l) serves in addition to the borehole heat exchanger as a heat source for the heat pump. This store is charged by the solar collector only.

2.1. System with integrated heat pump (system B)

In this system (see figure 2) the heat pump exclusively uses a borehole heat exchanger as heat source. The heat pump is fitted at the combistore (nominal volume 750 l) and the condenser is located in the auxiliary part of the combistore. The return flow of the space heating loop is connected to the combistore using a special device for stratified discharging of the store. The heat gained by the collector is added to the combistore via an internal heat exchanger which is also equipped with a special stratification device. An external heat exchanger is used for hot water preparation.

Results and discussion of testing laboratory model

Temperature control of the heater has been set at 33°C in the first experiment. Temperature in the up-flow pipe was greater than in the tank-accumulator during the first hour. When these temperatures became equal, the rising of temperature in tank-accumulator stopped at the level of 25°C (see Fig. 4). The duration of the cycles became longer (see Fig. 5), and heat accumulation was discontinued (see Fig. 6). Greater duration of cycles occurs when water in the accumulator is heated up completely and the heater fills only thermal losses. The difference in temperature between the descending pipe and the up-flow pipe was about 7°C. Only 0.6 liters of water flowed through heat exchanger during one cycle independently of heat source temperature.


The behavior of process was similar for other temperatures of the heater (see Fig. 7).

Time, h:mm


Time, h:mm

Fig. 7. Operating temperature in tank-accumulator for heater temperature of 33, 41, 49, 57 and



Подпись: Fig. 6. Operating capacity with respect to the time for a heater temperature of 33°C


Testing of laboratory model demonstrated stable operation and the heat flow depended on the capacity of the heater. This model of the circulating pump starts to act autonomously when the temperature difference between the descending pipe and the up-flow pipe exceeds 7 degrees. The proposed device can be used when heat has to transfer from a heat user source which is located below it, and can especially be used with a solar installation instead of an electrical circulating pump. The testing of a device which is integrated with a solar installation is currently underway.


[1] .Davidson J. H., Walker H. A., Lof G. O. G.. Experimental Study of a Self — Pumping Boiling Collector Solar Hot Water System. Journal of Solar Energy Engineering, 1989; 111 (3): 211 — 218.

[2] .Dobrianski Jury, Fieducik Jolanta Urz^dzenie do przekazywania ciepla w kierunku przeciwnym do konwekcji naturalnej. Patent PL 195490 B1 F24D 9/00 F03G 7/06

[3] .Dobriansky Y „Reverse thermosiphon’Undustrial heat engineering, Vol. 28, №6, 2006. — str. 44 — 48 (in Russian).

[4] .Dobriansky Yuriy. „Reverse thermosiphon”. IV International conference “Problems of Industrial Heat Engineering” Kiev, Ukraina 26 — 30 September, 2005. s 144.

[5] . Roberts C. C., Warrenville Jr. A Review of Heat Pipe Liquid Delivery Concepts/ Advances in heat pipe technology. London: Pergamon Press, Oxford, 1982: 693 — 702.

[6] . Walker H. A., Davidson J. H. Second — Law Analysis of a Two-Phase Self-Pumping Solar Water Heater. Journal of Solar Energy Engineering, 1992: 188 — 190.

[7] . Walker H. A., Davidson J. H.. Analysis and Simulation of a Two-Phase Self-Pumping Water Heater. Journal of Solar Energy Engineering, 1990 Vol. 112: 153 — 160.

The new solar combisystem concept


The complete solar combisystem consists of two units, the “Technical Unit” and the “Solar Store Unit” (see Fig 1 and Fig. 2).

In the technical unit all components like boiler, pumps, mixing valve, switching valves, heat exchangers, hot water preparation unit, expansion vessels, etc. are pre-installed. The main difference of this concept compared to existing ones, is the kind of integration of the condensing natural gas boiler. Due to the fact that the boiler is powerful enough for direct domestic hot water preparation, it can be avoided to heat the standby volume of the solar tank up to high temperatures for hot water preparation (typically 70°C or more). Therefore, the standby volume is only used at the temperature level needed for space heating operation, which results in much lower average temperature of the complete system and leading to higher overall performance of the heating system thanks to reduced heat losses of the tank and the pipes as well (detailed simulation results are presented in [2]). If hot water demand occurs and the temperature in the top of the solar tank is not high enough, the gas boiler immediately starts running in hot water preparation mode at high temperature level for direct hot water preparation in combination with the flat plate heat exchanger (see Fig. 3).


Fig. 3. Start/Stop frequency of the natural gas boiler during night for space heating; at about 06:30 and 07:42 domestic hot water preparation takes place (see also Fig. 2: Tc1-Tc20 in °C / DO1_Boil_S is the on/off

signal of the boiler).


Further advantage for the condensing natural gas boiler is the very low return temperature during domestic hot water preparation leading to higher condensation rate and higher efficiency. During periods where space heating load is less than the minimum power of the boiler which can be reached by modulation (5.7 kW) the use of the standby volume reduces significantly the start/stop frequency leading to less start/stop emissions and longer life time of the ignition unit in the boiler.

Due to this operation concept also the top of the solar tank (Tc1) never is heated to high temperatures by the boiler resulting in a higher heat storage capacity of the tank for the solar heating system.

Experimental Investigation

The preliminary experimental evaluation involved building and instrumenting an ISAHP system in a laboratory setting, based on the recommendations of component sizing given by Freeman [4].

The solar collector in Figure 1 was replaced with an auxiliary heater in order to perform controlled experiments. Quasi steady-state tests were run at range of constant input temperatures with all variables constant except for the natural convection flow rate. The natural convection flow rate varied throughout the duration of the test while the tank temperature increased. The preliminary results indicated that the original computer model over-predicted the actual COP of the system.

This discrepancy was determined to be due to an over-prediction of the heat exchanger effectiveness values for both the condenser and evaporator. After correcting the heat exchanger effectiveness values in the simulation, the results for power consumption and COP matched to within 3.0 % of each other for the 10oC test.

Resulting dimensions

Подпись:Подпись: primary energy savings [kWhprimary/a]The dimensioning method presented leads to a solar heating system with a comparably small collector area of 8.9 m2 and a storage device capacity of 0.67 m3; assuming a 60 kWh/m2 single family house located in Zurich with a hot water consumption of 3000 kWh/a. In Table 1 the optimal system configuration for these assumptions (called the base case) are summarised. To find the dimensioning parameters leading to the best cost/benefit ratio a number of simulation runs are necessary. The results of these runs — sorted by primary energy savings and

additional cost — are shown in Figure 2. Each dot in the chart represents a system with a different set of collector area and storage device capacity, leading to specific primary energy savings and additional cost. The quotient of these terms is the cost/benefit ratio which can be understood as the slope of a line through the origin meeting the respective point. The dimensions leading to the smallest gradient, which is also the tangent to a polynomial derived from all points, are the optimal dimensions (cf. [2]).

Load profiles

For the simulations daily and weekly load variations for the DHW with hourly resolution were created in order to separate between the ambient temperature dependant SH and independently varying DHW loads. To achieve realistic load profiles, they were created based on measurements from operational plants.

1.1. Measured load profiles

Measured hourly load data was obtained from six plants of various sizes. Measured monthly total loads were obtained from 18 plants of various sizes. It was found that the studied load profiles are basically of the same shape, i. e. mostly SH with a comparatively small summer loads (DHW and partly or mostly distribution losses), lower than 15% of peak load.

1.2. Load profiles for simulation studies

Based on the measured data representative generic load profiles were created. The DHW profile was created based on data when the SH is off or small. As the measured data included only the ambient temperature, no simulations regarding solar irradiation can be made for the same data set. Thus the SH load profile was created by matching visually the simulated to approximate the measured load characteristics while keeping the same peak load and approximate yearly total load. Case 1 represents the majority of the plants. As in the studied plant portfolio all cases except one had low summer load, a

Подпись: Fig. 1. Left: Hourly plot of the two simulated cases compared to a measured load profile. Right: Monthly load profile of the simulated cases compared to measured profiles. image156

fictive high summer load Case 2 was created by simply scaling up the DHW consumption of Case 1 and reducing the SH load keeping the same peak power. The studied climate data was TMY2 for Jokioinen (Finland), latitude 60.8°.

2. Simulations

SHW win was chosen as simulation tool because of ready and tested system models and it has also been used in a similar project earlier by TUG in Austria [1]. Yearly simulations were done with a sub­hourly resolution.

Boundary conditions


The simulation study is based on a single family house with a living area of 128 m2 located in Wurzburg, Germany. The roof area where the collectors are mounted is facing south with an inclination of 45o. The space heating demand of the building conforms to the current legal energy saving regulations (EnEV) and amounts to 71 kWh/ (m2 a) respectively 9090 kWh/a. The heating control is automatically adjusted to the outside temperature with a maximum flow/return temperature of 35/25°C. The heat demand for hot water amounts to 2945 kWh/a for a daily use of 200 litres at 45°C. The total heat demand (thermal requirement) for hot water preparation and space heating amounts to 12680 kWh/a, assuming heat losses of a conventional hot water store of 645 kWh/a. The flow temperatures for the brine were calculated based on measurements taken at a heat pump system using a borehole heat exchanger and range from 7°C in February to 20°C in July. The heat pump used in system A shows a COP (Coefficient of Performance) of 4,3 according to EN 14511 at 5K temperature difference. In system B a store-integrated heat pump is used which has a different thermal behaviour due to its positioning. This heat pump features a COP as per EN 14511 of 4,1 at 5 K temperature difference. The same flat plate collector with a total aperture area of 12 m2 and performance parameters of a “good” flat plate collector was used for both systems.

3. Results

Table 1 shows the most important results of the annual simulations carried out for the two systems A and B. In addition the thermal behaviour of a pure heat pump system without solar collectors (system C) was simulated for comparison. This system uses a similar combistore as system A, however with a smaller volume of only 400 litres. For all systems the electric energy consumption for the hydraulic pumps was not taken into consideration for reasons of simplification.

Table 1: Results of annual system simulation





„Usable hot water volume“





Collector gain: total/





thereof delivered in combistore




Heat losses of combistore





Delivered heat by heat pump





Seasonal performance factor [-]




electric energy consumption of the heat pump





Table 1 shows that system B requires the lowest amount of heat delivered by the heat pump in order to cover the entire heat demand for hot water preparing and space heating. Furthermore, this system provides the highest “usable hot water volume” and therefore offers the greatest hot water comfort. The positive thermal behaviour of system B is due to the efficient technology for hot water preparation (external heat exchanger combined with a controlled pump) and to the fact that at this system the highest solar energy gains are delivered to the combistore (3685 kWh). Moreover, the combistore of system B shows less heat loss than the combistore of system A. In system A the heat delivered from the collectors to the combistore is lower than in system B due to the control strategy of the collector loop pump: Within regular waiting periods during charging the buffer store it is checked if it is possible to charge the combistore. Nevertheless, system A requires the least electric energy consumption of the heat pump. That is because system A shows the highest COP due to the high temperature level in the buffer store which can occasionally be used as heat source for the heat pump. In comparison to system C (without solar thermal contribution) system A requires approx. 1000 kWh less electricity in order to operate the heat pump. The annual COP increases from 4,0 to 4,4 due to the solar thermal system. As can be expected, system A shows a high collector energy gain due to the additional buffer store.

4. Conclusions

The investigations have shown that the combination of a solar combisystem with a heat pump is a promising approach for saving primary energy. Due to the use of an additional buffer store in system A the electric energy consumption is less than in system B despite the higher delivered heat of the heat pump. This is due to the higher seasonal performance factor of the heat pump in system A. An

additional saving potential in system B can be utilised if the solar circuit is coupled with the brine circuit by a heat exchanger in order to preheat the return flow of the brine.

In order to compare different combinations of solar thermal systems and heat pumps in an objective way it is essential that standardised test — and evaluation procedures become available.


[1] Kuhl, L., Wendker, K., Fisch, N.: Praxistest von solarunterstutzten Warmepumpen-Heizsystemen; Tagungsband zum 17 Symposium Thermische Solarenergie, Otti-Technologie-Kolleg, Regensburg, Mai 2007; (2) ENV 12977-2: 2001: Thermal solar systems and components — Custom built systems — Part 2: Test Methods, ISBN 0-580-37754-7

[2] EN 14511-3: Air conditioners, liquid chilling packages and heat pumps with electrically driven compressors for space heating and cooling — Part 3: Test methods