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Energy and Buildings 100 (2015) 34–42

Contents lists available at ScienceDirect

Energy and Buildings

j o u r n a l h o m e p a g e : w w w . e l s e v i e r . c o m / l o c a t e / e n b u i l d

irect use of solar energy as heat source for a heat pump in omparison to a conventional parallel solar air heat pump system

. Lerch ∗, A. Heinz, R. Heimrath raz University of Technology, Institute of Thermal Engineering, Inffeldgasse 25/B, 8010 Graz, Austria

r t i c l e i n f o

rticle history: vailable online 7 March 2015

eywords: olar heat pump ir source heat pump ow energy building nglazed absorbers lazed collector

a b s t r a c t

In this paper different combinations of solar energy and heat pump systems are compared and inves- tigated through transient system simulations. The focus is on a small capacity heat pump (HP) with a heating capacity rate of 5.36 kW. The building (single family house 140 m2) which are considered in this paper have a space heating demand of about 45 kW h/m2 a. Based on a conventional air source HP and a parallel solar air source HP system, serial solutions for the combination of solar-HP systems are defined and analysed. For the consideration and evaluation of different systems through transient sys- tem simulations, a semi-physical HP-model is used, which offers the possibility to use two evaporators in series. With a combined parallel solar air source HP system the system performance (SPFSystem = 3.65) can be increased significantly compared to a conventional air source HP system (SPFSystem = 2.55). Unglazed

selectively coated absorbers as source for the HP offer the advantage, that the collector can be used as an air heat exchanger. If solar radiation is available, higher temperatures at the evaporator of the HP can be achieved, compared to a conventional air source HP system. By the integration of unglazed solar absorbers as source for the HP the HP performance can be increased significantly (SPFHP increases from 3.25 up to 3.55 compared to a conventional air source HP).

© 2015 Elsevier B.V. All rights reserved.

. Introduction

With regard to the continuous reduction of fossil energy esources and steadily increasing environmental pollution, solar nergy can be used to supplement space heating in buildings and lays an important role in reducing the dependence on conven- ional energy resources. In the Austrian building sector and also in ther European countries solar thermal systems and heat pumps re used increasingly to provide heat for space heating (SH) and or domestic hot water (DHW) preparation [1]. In most combined olar and heat pump systems, solar heat is used to charge a buffer torage, which is also charged by a heat pump (parallel system). owever, there are also systems, in which solar heat is used as heat

ource for the evaporator of the HP (serial system). Parallel systems ave the advantage of a lower complexity compared to serial sys- ems in terms of hydraulic connections and the system control and herefore they may be more robust and reliable [2]. In some con-

gurations of serial solar HP systems, unglazed absorbers are used s a heat source for the HP [3].

∗ Corresponding author. Tel.: +43 31 68737318. E-mail address: [email protected] (W. Lerch).

ttp://dx.doi.org/10.1016/j.enbuild.2015.03.006 378-7788/© 2015 Elsevier B.V. All rights reserved.

IEA SHC Task44 “Solar and Heat Pump Systems” was aimed at optimizing combinations of solar thermal energy and heat pump, primarily for one family houses (www.iea-shc.org/task44). The performance and relevance of combined systems using solar ther- mal and heat pumps were evaluated, a common definition of the performance of such systems and contributions to a successful market penetration of this combination of renewable technologies was provided. More than 100 systems from different participat- ing countries were surveyed during 2010 und 2011. The results of the survey showed that 70% of systems were found to be parallel solutions, 7% serial systems, 21% complex systems and very few regenerative solutions (unused solar energy in summer is used to regenerate boreholes in order to enable higher source temperatures for the HP). The number of air and ground source systems are about in the same range and some systems are operated only with solar collectors as source for the evaporator and many in several modes of operation [4]. An analysis of different collector types and system concepts showed that flat-plate collectors are used in nearly half of the systems (47%), whereas evacuated tube collectors are essential only in the fewest cases (2%). Instead, the choice between these two

types is frequently left open, i.e. affected by the conditions on-side as well as the preferences of client and installer (36%). Uncovered or unglazed absorbers are found repeatedly (7%), mainly in specific applications. Recently developed photovoltaic-thermal collectors

d Buildings 100 (2015) 34–42 35

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W. Lerch et al. / Energy an

re found only in few market-available solar HP-systems (8%) 5].

In combination with an ice storage system unglazed absorbers an be an attractive alternative to a ground source HP in case that, or example, regulations forbid to drill a borehole [2,6]. These serial ystems can also be seen as an alternative to air source systems hen efficiency or noise problems are of importance [7].

The aim of this paper is to investigate the seasonal perfor- ance of combined serial solar heat pump systems for a single

amily house, and to perform a comparison under the same oundary conditions to conventional air/water HP systems on one and and to parallel solar air/water HP systems on the other and. Within Task44 common boundary conditions for the DHW emand and the building heat load and performance indicators ave been defined and implemented [8], which are used within this ork.

. Solar thermal and heat pump systems

In the national Austrian Project “SolPumpEff” [9] 6 different olar- air source HP systems are considered in terms to the system fficiency. These systems have been defined within the framework f the project and shown in this paper.

.1. System A

Fig. 1 shows the hydraulic layout of System A, which is an ir/water HP system without any solar thermal system. The HP rovides heat not only to the buffer storage but also directly to he heating system of the building. The buffer storage is divided nto a DHW and a SH volume. The storage volume available for HW is based on the assumed daily hot water demand (0.2 m3). he total buffer storage volume (DHW and SH) is 0.3 m3. The stor- ge volume results from optimization calculations. As backup an lectrical heater is placed in the buffer storage (space heating vol- me), which provides heat at times when the heating capacity rate

f the air/water HP is not sufficient to cover the heating demand. he domestic hot water (DHW) preparation is done via an external eat exchanger.

Fig. 1. System A.

Fig. 2. System B.

2.2. System B

In System B a solar system is installed in a parallel way, whereby 14 m2 of covered flat plate collectors with selectively coated absorbers are used. Compared to System A the buffer stor- age volume is increased to 1 m3, in order to have enough volume for the solar collectors. In Fig. 2 the integration of the solar ther- mal system is shown. The solar loop is divided into a brine (solar side) and a water loop (buffer storage), whereby the heat transfer between these takes place via a plate heat exchanger. Solar energy is only used to charge the buffer storage. The air/water HP system is installed in a parallel way, charging only the upper part of the store (0.4 m3).

2.3. System C

As shown in Fig. 3, System C is a serial solar HP system. As heat source for the HP only unglazed, selectively coated absorbers are used. Solar energy can either be used to charge the buffer storage directly or as heat source for the HP, a simultaneous operation is not possible. In times of little or no solar irradiation heat from the ambient air can be collected by the uncovered absorbers and used by the HP.

2.4. System D

System D is similar to System C with the difference of an addi- tional ice storage (Fig. 4), which is used as source for the evaporator of the HP. As in System C unglazed, selectively coated absorbers are used. The buffer storage (hot water storage) is charged with

Fig. 3. System C.

36 W. Lerch et al. / Energy and Buildings 100 (2015) 34–42

p p e I c

2

e t o t t t i e o

2

t e

Fig. 4. System D.

riority. If the HP is operated and the solar thermal system sup- lies energy to the ice storage at the same time, then energy for the vaporator of the HP comes directly from the solar thermal system. n this case higher source temperatures than from the ice storage an be achieved at the evaporator of the HP.

.5. System E

Fig. 5 shows a solar air/water HP system, in which the solar nergy is used to charge the buffer storage on the one hand, and o preheat the air at the outdoor unit of the air/water HP on the ther hand. For the solar collectors, glazed collectors with selec- ively coated absorbers are used. If the HP has to provide heat and he collector temperature level is higher than the temperature of he ambient air, solar heat is used to preheat the ambient air before t enters the evaporator of the HP. For this purpose a brine/air heat xchanger, which is connected to the solar loop, is installed in front f the evaporator.

.6. System F

Also in System F solar heat is used to achieve higher evaporation emperatures (Fig. 6). The difference to System E is, that the heat xchanger is directly integrated into the refrigerant cycle of the HP

Fig. 5. System E.

Fig. 6. System F.

after the air source evaporator (in refrigerant flow direction). Addi- tional solar heat supply to the HP is activated, when the collector temperature is higher than the ambient air temperature. Glazed collectors with selectively coated absorbers are used.

3. Boundary conditions

3.1. Building

For the system simulations the boundary conditions of the IEA SHC Task44 are used [10,11]. In Task44 three different buildings concerning the space heating demand (single family house, heated floor area 140 m2) have been defined. For the climate Graz (aver- age climate data from 2001 to 2010, average temperature 10.65 ◦C, global radiation 1196 kW h/a, diffuse radiation 605 kW h/a, heating degree days 3102 HDD(20 ◦C/12 ◦C))

1 [12] the buildings have a spe- cific heat demand of 15, 45 and 100 kW h/m2 a. In this paper the simulation results only for the SFH45 are shown. For the SFH45 a floor heating system is assumed. The flow temperature is controlled depending on the ambient air temperature (for SFH45 35/30 ◦C at the design ambient temperature, radiator exponent 1.1).

The heat load of the building for the location Graz was used for the dimensioning of the HP for a specific heat demand of 45 kW h/m2 a. For the transient simulation of the heating system the building was considered by means of a transient thermal build- ing simulation in TRNSYS [13], with the building model defined in [19].

The heat load of the building has been calculated in advance for the design ambient temperature of −12 ◦C in order to be able to size the heat pump. The HP is sized in a way to cover the design heat load of the building (4.86 kW) increased by 0.5 kW (addition for DHW preparation) at the operation point A2/W35 (5.36 kW). The operation point A2/W35 is selected on the basis of a bivalent heat preparation system. As a second heat source an electrical heater in the water storage is installed for times, where the heating capacity rate of the HP is lower than the demand. With this dimensioning less than 5% of the total heat demand have to be covered by the electrical heater with the used climatic conditions. The advantages of a smaller HP dimensioning are fewer starting cycles of the HP

and thus reduced losses and lower investment costs.

1 Based on monthly data (average monthly data of the years 2001–2010) hourly values of all parameters are calculated with the Software METEONORM [12], which uses a stochastic model. The resulting time series represents a “typical yearät the considered location.

W. Lerch et al. / Energy and Buildings 100 (2015) 34–42 37

Table 1 TRNSYS models used within this work.

Type Description Documentation Description/parameters

Type 877 Heat pump model [14] Air HP, power: SFH45, 5.36 kW, (COP at A2/W35 = 4.22) Type 843 Ice storage model [15] Ice storage: internal heat exchanger (pipe diameter 20 mm, distance between pipes

50 mm) Type 832 Collector model (glazed) [16] Model glazed collector; �0 = 0.8, a1 = 3.5 W/(m2 K), a2 = 0.015 W/(m2 K2 ), slope: 45◦ ,

orientation: south, 14 m2

Type 203 Collector model (unglazed) [17] Model unglazed collector with condensation; �0 = 0.954, a1 = 9 W/(m2 K), a2 = 0 W/(m2 K2 ), slope: 45◦ , orientation: south, 30 m2

Buffer 3 3

Stand SFH45

3

T i s h (

3

w o a i w t

3

a S p p i U

t u c t r e u t t o e i ( s d

3

o a g d h i

Type 340 Storage model [18] Type 56 Building model [19] Type 362 Radiator model [20]

.2. Domestic hot water

The domestic hot water demand is used as defined in Task44. he DHW preparation is done via a plate heat exchanger, which s connected to the buffer storage, in all systems. The hot water et temperature is defined with 45 ◦C. From the annual DHW eat demand results an average daily heat demand of 5.85 kW h/d 2076 kW h/a).

.3. Simulation models

The transient system simulations are performed using the soft- are TRNSYS [13], which is well suited for the transient simulation

f buildings and HVAC systems, as many component models are vailable either in the standard library or from different researchers n the TRNSYS community. The main component models (types),

hich are used for the simulations within this work, are listed ogether with the most important parameter settings in Table 1.

.3.1. Heat pump model The basis for the compression heat pump model Type 877 is

n EES-model, that has originally been developed by NTB Buchs in witzerland [21] using the ARI model for the simulation of the com- ressor performance [22]. Based on this model a TRNSYS model was rogrammed at Institut für Solartechnik SPF and further developed

n a cooperation between the Institute of Thermal Engineering, Graz niversity of Technology and SPF [14].

Type 877 is a semi-physical model based on a calculation of the hermodynamic refrigerant cycle and the thermal properties of the sed refrigerant. A performance map of the compressor is used for alculating the compressor efficiency and the electricity consump- ion. Fig. 7 left shows a schematic view of the HP cycle and Fig. 7 ight an example process in the T/h-diagram with all active heat xchanger. Compared to performance map models, which are often sed for annual simulations of heat pump systems, the model has he advantage that it can be used to study the effect of changes in he HP circuit, like e.q. a different refrigerant, varying mass flows n the source and sink side or changes of the UA-value of the heat xchangers. It is also possible to implement a second evaporator n series—one for air-source and the second for a solar heat source brine), or to use an additional desuperheater on the high pressure ide. A more detailed description of a comparison with measured ata (validation) can be found in [23,24].

.3.2. Ice storage model For the simulation of the ice storage in System D, type 843, devel-

ped at the Institute of Thermal Engineering, was used. This model llows the simulation of a cuboid-shaped ice storage with an inte-

rated pipe heat exchanger. It is very important to consider that uring discharging an ice layer forms on the outside surface of the eat exchanger pipes, when the temperature drops below the freez-

ng point. This layer of ice acts like an insulation layer and reduces

storage; volume: with solar 1 m , without solar 0.3 m ard type; multi zone model : flow-/return temperature = 35/30 ◦ C

the heat transfer between the liquid water in the tank and the heat transfer fluid in the heat exchanger pipes depending on the ice layer thickness. Ice formation and growth on the tubes and the increasing thermal resistance of the ice is considered by means of a finite dif- ference formulation of the transient Fourier equation in cylindrical coordinates. For the modelling of the phase change between solid and liquid the enthalpy method is used [25,26]. As a simplification the temperature of the liquid water in the storage is assumed to be the same in the whole tank.

The heat exchanger is assumed as a pipe, which is laid in loops inside the storage. The pipe diameter, the wall thickness, the pipe spacing distance and the number of the parallel cycles can be cho- sen as parameters. The heat exchange between the heat carrier fluid and the inner wall of the heat exchanger pipes is calculated using an empirical model for the flow in cylindrical pipes [27]. The heat losses/gains of the storage to/from the ambient are calculated through a heat loss coefficient UALoss [W/K] which is provided as an input.

3.3.3. Unglazed collector model [17,28] The thermal behaviour of the collector depends on a character-

istic curve under steady state conditions according to EN 12975-2 [29] for testing of the performance of unglazed solar collectors.

It has to be mentioned that for the unglazed absorbers not all physical effects were taken into account in this paper. As example: the loss of selectivity of the surface when the unglazed collector is covered by water droplets from condensation of moisture from the ambient air [7,30].

3.4. Control strategy, sensor positioning

In this section the control strategies for the different systems are described. The operating states of the solar circle, the HP and the electrical heater depending on different temperature readings are shown in Table 2. In Fig. 2 the reference system for the parallel solar thermal HP system is shown.

The positions of the different temperature sensors for the con- trollers are shown in Figs. 1 and 2. The temperature Theat circuit corresponds the required flow temperature of the space heat- ing system. The flow temperature of the heating system changes depending on the ambient air temperature.

All systems are optimized to get the highest system efficiency. To get the right system configuration and control strategy the storage volumes, the inlet and outlet heights of the storage, the positions of the temperature sensors and also the control temperatures are optimized.

4. Performance figures

The performance indicators are used in this work as defined in IEA SHC Task44 [31]. The most important indicators are the

38 W. Lerch et al. / Energy and Buildings 100 (2015) 34–42

r2r5r8

r9 r1r10

d2d5w5w8

a10a9 b1b10

evapo rator air

evapo rator brine

retaehrepusedresnednoc

compress orexpan sion valve

r7

r8

r2

r4=r5=r6 r3 d2

d4=d5

d3 w5=w6

w7w8

r1

r11r10r9

a9

a10 b1

b10

T

exam

s p

f m a

S

S

S

d P P i

T C

Fig. 7. Left: schematic of the heat pump cycle; right: schematic

easonal performance factor of the system (SPFSystem), the seasonal erformance factor of the HP (SPFHP) and the solar fraction (SF).

The SPF of the entire system with penalties (if required energy or DHW and SH cannot be covered through the heating system, the

issing part of energy is also considered in the system performance nd is called penalty) is defined as:

PFSystem = ∫ (

Q̇SH + Q̇DHW )

× dt ∫ (

Pel,tot + Pel,SH,pen + Pel,DHW,pen )

× dt (1)

The definition of the SPF of the HP is:

PFHP = ∫

Q̇HP,Cond × dt∫ Pel,HP,tot × dt

(2)

The solar fraction as:

F = 1 − ∫ (

Q̇HP,Cond + Pel.heater )

× dt ∫ (

Q̇SH + Q̇DHW )

× dt (3)

Q̇SH is the heat power space heat; Q̇DHW is the heat power ˙

omestic hot water; QHP,Cond is the heat power HP condenser;

el,SH,pen is the electricity consumption penalty space heating; el,DHW,pen is the electricity consumption penalty DHW; Pel,SH,pen s the electricity consumption penalty space heating; Pel,HP,tot is

able 2 ontrol strategy.

Action State Condition

Solar cycle: charging the water storage

Pumppri on Tcoll ≥ TS Solar + 7 K Pumppri off Tcoll ≤ TS Solar + 3 K Pumpsec on THX ≥ TS Solar + 7 K Pumpsec off THX ≤ TS Solar + 3 K

Solar cycle: charging the ice storage (System D)

Pumppri on Tcoll ≥ Tice storage + 7 K AND Tcoll ≤ TS Solar

Pumppri off Tcoll ≤ Tice storage + 3 K OR Tcoll ≥ TS Solar

Solar air preheating (System E and F)

Pumpsolar HP on Tcoll ≥ Tambient air + 5 K Pumpsolar HP off Tcoll ≤ Tambient air + 3 K

Charging of the domestic hot water volume

DHW loading on THW ≤ 54 ◦ C DHW loading off THW ≥ 57 ◦ C

Charging of the space heating volume

SH loading on TSH ≤ Theat circuit −3 K SH loading off THP entry ≥ Theat circuit

Electrical heater Electrical heater on TSH ≤ Theat circuit−5 K Electrical heater off TSH ≥ Theat circuit−3 K

h

ple process in the T/h-diagram with all heat exchangers active.

the electricity consumption of entire HP (compressor, fan, con- troller); Pel,tot is the electricity consumption of all components (HP, controller, pumps, electrical heater); Pel,heater is the electricity con- sumption electrical heater.

5. Results

5.1. System comparison

The different systems have been evaluated by means of the above described performance indicators. The results are shown in Table 3. To cover the DHW demand an energy demand of 2076 kW h/a, and for the SH demand 6385 kW h/a are used. The comparison of Systems A and B shows that by integration of a solar thermal system (14 m2) SPFSystem can be increased from 2.55 to 3.65 with the used assumptions and boundary conditions. The total elec- tricity consumption of the system is reduced by about 30%. SPFHP is reduced, as the solar collectors are providing heat mainly dur- ing the warm season, when the HP would otherwise be operated with a higher efficiency. However, SPFSystem and Pel,tot are the more important figures from an overall system point of view.

Fig. 8 shows additional to Table 3 exemplary results for four dif- ferent heating systems (A–D) with different configurations. With a conventional air/water HP system (A) a SPFSystem of 2.55 is achieved. With a conventional parallel solar and air/water HP system (B), with

2.0

2.5

3.0

3.5

4.0

4.5

10 20 30 40 50

SP Fs

ys te

m

coll ector area ung lazed [m² ]

1.5 m³ 0.9 m³ System D 0.6 m³ 0 m³ System C

air/water HP ( with 14 m² glazed flat plate collector (parallel, System B)

air/wate r HP (withou t sol ar, System A)

ice storage volum e

Fig. 8. SPFSystem for System A and B and different configurations of System C and D.

W. Lerch et al. / Energy and Buildings 100 (2015) 34–42 39

Table 3 Simulation results.

System (−) Collector area (m2 )

Buffer storage volume (m3 )

Ice storage volume (m3 )

SPFSystem (−) SPFHP (−) SF (%) Pel,tot (kW h) Pel,heater (kW h)

A – 0.3 – 2.55 3.25 – 3327 274 B 14 1 – 3.65 3.04 31.0 2317 285 C 30 1 – 3.53 3.37 26.2 2401 485

3 3 3

a 3 H a w b a t i e

t b a p a m o t t s t h S s

5 a

p a

F u

D 30 1 0.6 E 14 1 – F 14 1 –

collector area of 14 m2 (covered flat plate collectors) a SPFSystem of .65 can be reached. Additionally SPFSystem for combined solar and P systems with unglazed absorbers for different collector areas nd ice storage volumes is shown. To reach the same SPFSystem as ith the parallel HP system with glazed collectors (14 m2), it would

e necessary to use an unglazed absorber area of 30 m2 (no ice stor- ge). With about 12 m2 of these collectors it is still possible to reach he same SPFSystem as with a conventional air/water HP system. The ntegration of an ice storage further increases the system efficiency, specially with a small collector area.

The simulation results for System E show that by preheating he ambient air the total electricity consumption of the system can e reduced by 1–2% compared to System B with the used bound- ry conditions and assumptions. The increased performance can be rimarily explained by an increased performance of the HP (higher ir temperatures at the evaporator), but the efficiency improve- ent potential is rather small. This is due to the fact that the use

f solar heat at the evaporator of the HP is only beneficial, when he radiation on the collector field is low. At times when the radia- ion exceeds a certain limit, usually a direct charging of the buffer tore by the collectors is favorable [32]. The low overall benefit by his kind of operation can be explained by the low amount of solar eat, which is occurring below this limit. The comparison between ystem E and F does not show a big advantage for one of the two ystems (compare with Table 3).

.2. Comparison of the evaporator inlet temperature for system C nd D

Fig. 9 shows cumulative diagrams, in which the ambient air tem- erature and the evaporator inlet- and outlet temperature is shown s a function of the supplied heat at the evaporator of the HP. With

ig. 9. Heat provided at the evaporator of the HP with different evaporator inlet- and outl nglazed collector, 0.6 m3 ice storage.

.56 3.55 23.8 2385 492

.68 3.10 30.4 2298 275

.70 3.11 30.4 2289 273

System C (Fig. 9, left) about 4500 kW h are supplied by the unglazed absorbers.

With System C about 900 kW h are provided to the evaporator with a brine inlet temperature higher than the ambient air tem- perature. This occurs, when solar radiation is available, which lifts the brine temperature in the collector to a level higher than the ambient air. With System D (Fig. 9 right) 4800 kW h are supplied by the unglazed absorbers and the ice storage of the solar system. The phase change (liquid/solid) in the ice storage can be seen in the course of the evaporator inlet temperature, when compared to System C. A comparison shows, that the integration of an ice storage causes – on average – an increase of the evaporator inlet temperature, which results in a higher efficiency of the brine/water HP (SPFHP). The system efficiency (SPFSystem) increases only slightly with the chosen boundary conditions. This is due to a lower amount of solar energy supplied directly to the buffer storage.

5.3. Monthly simulation results

In Figs. 10–12, monthly heat balances for Systems A–F are shown. The left bars in the figure show the energy flows into the sys- tem and the bars on the right side of each month show the energy consumption. Additionally average monthly performance figures are shown (SPFSystem, SPFHP and SF).

On the left side of Fig. 10 the simulation results for the air HP sys- tem (System A) are shown. On the Input side of the energy flow the air source for the HP, the electricity consumption of the compressor of the HP and the electricity consumption of the electrical heater are shown, on the consumption side the energy for the DHW and

SH demand and also the heat losses of the storage, the HP and the pipes. The average monthly SPFHP decreases in the summer months, as the HP is operated only in DHW mode, where it is less efficient than in SH mode. The right sight of Fig. 10 shows the results for

et-temperatures; left: System C, 30 m2 unglazed absorbers; Right: System D, 30 m2

40 W. Lerch et al. / Energy and Buildings 100 (2015) 34–42

Q solar storage Q solar HP Q air Wel comp HP Wel hea ter Q DHW Q SH Q heatloss st orage s, HP Q heatloss pipes SPF System SPF HP solar frac�on

0

500

1000

1500

2000

2500

Jan Feb Mär Apr Mai guAluJnuJ Sep Okt Nov Dez

[k W

h]

0

1

2

3

4

5

6

7

8

9

10

Jan Feb Mär Ap r Mai Ju n Jul Aug Sep Okt Nov Dez SP F

Sy st

em ,a

ve ra

ge m

on th

ly S

PF H

P, s

ol ar

fr ac

� on

Fig. 10. Monthly heat balances, left: System A, air/water HP. Right: System B, air/water HP, 14 m2 glazed flat plate collectors.

Q solar storage Q solar HP Q air Wel comp HP Wel hea ter Q DHW Q SH Q heatloss st orage s, HP Q heatloss pipes SPF System SPF HP solar frac�on

0

500

1000

1500

2000

2500

Ja n Fe b Mär Apr Ma i J un Jul Aug Se p Okt Nov Dez

[k W

h]

0

1

2

3

4

5

6

7

8

9

10

Jan Feb Mä r Apr Mai Jun Jul Aug Sep O kt Nov Dez S PF

S ys

te m

, a ve

ra ge

m on

th ly

S PF

, so

la r

fr ac

� on

Fig. 11. Monthly heat balances, left: System C, HP and 30 m2 unglazed flat plate collector. Right: System D, HP and 30 m2 unglazed flat plate collector, 0.6 m3 ice storage.

Q solar storage Q solar HP Q air Wel comp HP Wel heater Q DHW Q SH Q heatloss storages, HP Q heatloss pipes SPF System SPF HP solar frac�on

0

500

1000

1500

2000

2500

Ja n Feb Mä r Apr Ma i J un Jul Aug Sep Ok t Nov Dez

[k W

h]

0

1

2

3

4

5

6

7

8

9

10

Jan F eb Mär Apr Mai Ju n Jul Aug Sep Okt Nov Dez

SP F

Sy st

em , a

ve ra

ge m

on th

ly S

PF H

P, s

ol ar

fr ac

� on

Fig. 12. Monthly heat balances, left: System E, 14 m2 glazed flat plate collector. Right: System F, 14 m2 glazed flat plate collector.

d Bui

S D f t s

H s a t M c v I t t

( e i s t t t u p H

6

p u A c 1

s o a ( f a h e s

t h s c t s e t w i e p e t w s C i

W. Lerch et al. / Energy an

ystem B with 14 m2 glazed selective coated collectors (parallel). ue to the integration of the solar thermal system the system per-

ormance can be significantly increased. From May to September he total heat demand can be covered via the solar thermal ystem.

In Fig. 11 the simulation results for Systems C and D are shown. ere the solar energy which is directly used to charge the buffer

torage is shown separately from the solar energy which is used s heat source for the HP (or the ice storage). In Systems C and D he unglazed absorbers are used as heat source for the HP. From

ay to September in System C solar heat is only used to directly harge the buffer storage, because the heat demand can be covered ia the solar thermal system and there is no operation of the HP. n System D solar heat is used to charge the buffer storage but also he ice storage. In the summer months there is also no operation of he HP.

Fig. 12 shows the results for Systems E and F. On the left side System E) the ambient air is preheated through an external heat xchanger. The difference in System F is that, that the solar cycle is ntegrated in the refrigerant loop. On the input side of the monthly imulation results the solar heat, which is used to directly charge he buffer storage and that which is used to increase the evaporator emperature is shown. The solar energy which is used to preheat he air entering at the outdoor unit (System E) and also which is sed in a second evaporator (System F) is quite low. As a result the erformance increase compared to a parallel-only solar air/water P system is relatively poor.

. Summary and conclusions

In this work parallel and serial solar HP systems for heating urposes in buildings are evaluated through transient system sim- lations and compared by means of defined performance figures. s a basis for the comparison a low energy building with a spe- ific heat demand of about 45 kW h/(m2 a) and a gross floor area of 40 m2 is used.

As a reference for the comparison, a conventional air/water HP ystem is used, which reaches a SPFSystem of 2.55. The integration f 14 m2 of glazed solar thermal collectors with selectively coated bsorbers, which charge the buffer storage in parallel to the HP parallel system), results in a significant increase of the system per- ormance to 3.65. The total electricity consumption is reduced by bout 30%. SPFHP is reduced, as the solar collectors are providing eat at times, when the HP would otherwise be operated with a high fficiency. However, the more important figures from an overall ystem point of view are SPFSystem and Pel,tot.

As an alternative to the parallel system four different possibili- ies of serial systems were defined, where solar heat is also used as eat source at the evaporator of the HP. One system uses unglazed electively coated absorbers, with the possibility to either directly harge the buffer store or to be used as the single heat source of he HP. The simulation results show, that in order to reach the ame SPFSystem as with the parallel HP system, it would be nec- ssary to use about 30 m2 of unglazed absorbers. With 12 m2 of hese collectors it is still possible to reach the same SPFSystem as ith the conventional air/water HP system. When an ice storage

s integrated into the system, SPFSystem can be further increased, specially with small collector areas. However, the benefit com- ared to a system without ice storage is rather low and from on conomic point of view it might be better to invest the costs of

he ice storage into a larger collector area. Such systems – with or ithout ice storage – could be an interesting alternative to ground

ource HPs, when for example regulations forbid to drill a borehole. onsidering noise problems or efficiency issues they could also be

nteresting as a replacement for air source systems.

[

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ldings 100 (2015) 34–42 41

In one of the investigated systems, solar heat from glazed collec- tors is used both to charge the buffer storage and to preheat the air entering at the outdoor unit of the air/water HP. In an alternative configuration solar heat is charged directly into the refrigerant cycle via an additional evaporator connected to the solar loop, which is installed in series to the air source evaporator. The results show that the total electricity consumption of the system can be reduced by 1–2% compared to the parallel solution under the used boundary conditions. The increased performance can be primarily explained by the higher SPFHP caused by higher evaporation temperatures. The relatively small efficiency improvement is due to the fact that the amount of solar heat which is to use in a beneficial way at the evaporator of the HP is small. Due to the low efficiency improve- ment these systems do not appear as an interesting alternative to a simple parallel solution considering the increased system com- plexity concerning the hydraulics and the control.

It can be summarized, that the combination of HP and solar ther- mal systems can be an efficient and attractive solution for heating and DHW preparation in low energy buildings. The most simple solution for this combination analyzed in this work is a parallel sys- tem, where both, the air source HP and the solar collectors charge a buffer storage. Serial solutions, where solar heat is also used as additional heat source for the HP can only slightly outperform the parallel system. Systems with unglazed absorbers as the only heat source of the HP can be an interesting alternative to both, ground- and air-source HPs.

Acknowledgements

This work was financed through the research- and technology program “Neue Energien 2020” by the Climate and Energy Fund in Austria (Grant number: 825546).

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