Novel speed-controlled exhaust-air to supply-air heat pump combined with a ventilation system

Applied Thermal Engineering 162 (2019) 114230

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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng

Research Paper

Novel speed-controlled exhaust-air to supply-air heat pump combined with a ventilation system

T

Dietmar Siegele , Fabian Ochs, Wolfgang Feist ⁎

Unit for Energy Efficient Building, University of Innsbruck, Technikerstraße 13, AT-6020 Innsbruck, Austria

HIGHLIGHTS

supply air heat pump in combination with a ventilation system. • Innovative numerical model for full range of boundary and operation conditions. • Validated • An overall system performance between 2.5 and 4.5 can be expected. ARTICLE INFO

ABSTRACT

Keywords: Exhaust-air heat pump Speed-controlled air-to-air heat pump Ventilation system

This paper presents the detailed steady-state measurement results of a novel speed-controlled exhaust-air to supply-air heat pump combined with a ventilation system. Compared with conventional systems, the heating power can be more than doubled to approximately 2.5 kW using recirculation air. Therefore, such cost-effective systems can be utilized not only in high-energy-efficient buildings but also in buildings with higher heating loads, such as in case of renovations. A functional model was developed and tested in the laboratory. The measurement results demonstrated an overall system performance of over 4.5 for minimal heating power at +10 °C and 2.5 for maximum heating power at −7 °C. A simplified physical model for refrigerant cycle was presented and validated. It provided results with a measurement accuracy of 8%. The model was used to demonstrate further optimization potential. According to these results, a 5% improved system performance can be achieved by increasing the condenser size.

1. Introduction Improving the quality of a building stock considerably lowers the global energy consumption. Appropriately designed ventilation systems have been demonstrated to reduce heating and cooling energy demand [1,2] and improve thermal comfort and indoor air quality [3,4]. During winter conditions, the enthalpy of exhaust air in a ventilation system with a heat or enthalpy exchanger is higher than that of ambient air and can therefore be used as a source for a heat pump. Such so-called exhaust-air heat pump uses the air after the air-to-air heat exchanger in a ventilation unit. In contrast, an extract-air heat pump (also called exhaust-air heat pump in the literature) uses room air as source and replaces the air-to-air heat exchanger. A classification of such systems was made in [5]. Compared with an extract-air heat pump, the temperature level of the evaporator in an exhaust-air heat pump is lower but the air is almost saturated. However, the enthalpy is still higher than that in ambient air [6].

⁎

For high-energy-efficient buildings, such systems can supply sufficient power to satisfy the designed heating load using the supply air. For deep-impact renovations and passive-house quality, [7] theoretically demonstrated that such systems could achieve the highest overall performance. They were compared with other systems such as the extract-air heat pumps without heat recovery. In [8], an experimental comparison of four types of ventilation heat-recovery systems confirmed this result. Using the in-situ monitoring study in [9], this concept was proven for a renovation case. Nevertheless, such systems suffer from some disadvantages and faces challenges. For instance, the maximum heating power that these systems can provide is limited by the supply-air flow rate [10]. Moreover, the volume flow should not exceed the hygienic-air circulation because of the increased risk of dry room air during winter period even if used in combination with an enthalpy exchanger [11]. This constraint limits the effective specific heat load that can be provided by the supply air to approximately 10 W/m2. Thus, such systems are not applicable in

Corresponding author. E-mail address: [email protected] (D. Siegele).

https://doi.org/10.1016/j.applthermaleng.2019.114230 Received 7 June 2019; Received in revised form 19 July 2019; Accepted 8 August 2019 Available online 09 August 2019 1359-4311/ © 2019 Elsevier Ltd. All rights reserved.

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Nomenclature

pressure ratio [–]/ = pHP /pLP relative humidity [%] correlation factor two-phase region [–] humidity ratio [kgwater/kgdryair]

Latin letters

A Co cp d dh Dc G h jH L m N n Nu p Pel Pr pr Q Re s St Sl t T UT UA V w x

area of heat exchanger [m2] convection number [–] specific heat capacity for constant pressure [J/(kg·K)] thickness [m] hydraulic diameter [m] displacement of compressor [m3] mass flux [kg/(m2·s)] enthalpy [J/(kg·K)] Colburn factor [–] characteristic length [m] mass flow [kg/s] compressor speed [rpm] no. [–] Nusselt number [–] pressure [Pa] electrical power [W] Prandtl number reduced pressure [–]/ pr = p / pc thermal power [W] Reynolds number [–] fin distance [m] distance of tubes in vertical direction [m] distance of tubes in horizontal direction [m] fin thickness [m] absolute temperature [K] overall heat-transfer coefficient [W/(m2·K)] thermal-loss coefficient [W/K] volume flow [m3/h] velocity [m/s] mass fraction in two-phase region [–]

Indices

air abs amb cond comp exh ext evap fin G h HG HG* HP i in inv ise L loss LP o out pipes room ref rec set SH suc sup TP tube vent vol w '

Greek letters

hv p

µ

heat-transfer coefficient [W/(m2·K)] latent heat of evaporation [J/kg] differential pressure [Pa] efficiency dry-bulb temperature [°C] thermal conductivity [W/(m·K)] dynamic viscosity [Pa·s] density [kg/m3]

case of renovations in which specific heat loads of 20–25 W/m2 can be demanded [12]. One possibility of increasing the heating power is to use recirculated secondary air. This recirculation air can be mixed with the supply air or separately distributed. The second option allows a better distribution of the heating load, e.g., by controlling the volume flow of the recirculation air. Moreover in addition to the supply air rooms the corridor (where the recirculation air is taken from) is also heated, which leads to an increase in thermal comfort [13]. The increased volume flow additionally enables such systems to be used for cooling purposes. However, to increase the total heating power of the heat pump, additional ambient air (independent of the fresh air from the ventilation unit) has to be used, in addition to the exhaust air. Two heat pumps with two different total heating power capacities are then developed. The focus of this paper is to present this new concept and show the range of possible operating modes of such solution using laboratory measurements in a test rig under reproducible conditions. A physical calculation model is validated and used to show the optimization

air absolute ambient air condenser compressor exhaust air extract air evaporator fins gas humidity (measurement point) hot gas (compressor outlet) hot gas (compressor outlet, theoretical point) high pressure at compressor outlet inner inlet inverter isentropic liquid losses low pressure at compressor inlet outer outlet pipes installation room refrigerant recirculation air set point superheating suction (compressor inlet) supply air two phase tube ventilation volumetric water saturation

potential of such systems. In contrast to previous studies, we show how exhaust-air heat pumps can deliver higher heating power independent of the necessary hygienic air flow. Moreover, no measurement data of such systems are available in the literature, and in the present study, the high-quality measurement results of an exhaust-air heat pump with a speed-controlled compressor are presented, which can be used for further investigations. 2. System description Fig. 1 shows the –h diagram of the refrigerant cycle (solid line) of an exhaust-air heat pump with two condensers. The air temperatures are indicated by dashed lines. Compared with other systems, the temperature differences between the air inlet and outlet are much higher. This result explains why the pressure drops in the refrigerant cycle (such as in the evaporator, see dotted line) have no significant effect, and a relatively high subcooling can be reached. The reason is that the thermal conductance in the refrigerant side in a two-phase region is 2

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and inverter. The exhaust air flows over the upper part of the evaporator. A minimum volume flow of 35 m3/h of ambient air is maintained during the heat-pump operation to cool the compressor and inverter. If necessary, additional ambient air can be used to increase the heating power. An electrical expansion valve (EEV) is used to control the superheating of the evaporator to approximately 5 K. The evaporator is a plate-fin-tube heat exchanger (PFTHE) and has a condensate sump. The condensate directly flows out of the unit through a drainage system. The specifications of all components, refrigerant pipes, and built-in parts are listed in Table 2 of Appendix A. 3. Method 3.1. Refrigerant-cycle model Fig. 1. –h diagram of an exhaust-air heat pump with two condensers without losses.

The model considers the evaporator, condensers, compressor, and expansion valve. To account for the refrigerant charge, the thermal losses of the pipes and built-in parts are also modeled. Pressure drops are only considered in the evaporator and condensers. Eq. (1) shows the energy balance of the refrigerant cycle, and Eqs. (2) and (3) express the coefficient of performance (COP) of the heat pump and the overall system, including that of the ventilation. Eqs. (4)–(8) show the calculation of the power used for the analysis of the measurement and simulations. In the presented system configuration the heat losses of compressor and inverter can be used by the evaporator [compare Fig. 2 and Eq. (6)]. The heat losses of the pipes, expressed in Eq. (7), are shown for the hot-gas refrigerant pipe, and a similar equation expresses that for the subcooling refrigerant pipe.

significantly higher (5–15 times in the evaporator) than that in the air side [14]. Thermal conductance comparable with that in the air side can be obtained only in the liquid and gas phase of the refrigerant. The two variable-speed heat pumps with nominal total heating power capacities of approximately 1.4 and 2.5 kW use the enthalpy of the exhaust air of a mechanical ventilation system with heat and moisture recovery. The devices consist of indoor and outdoor units that are flexible with respect to the installation. Fig. 2 shows a scheme of the system with the air flow and refrigerant cycle. In the primary mode of operation, both condensers are used. In this mode, the superheated vapor flows into the condenser of the recirculation air and subsequently through the condenser of the supply air. In the second mode of operation, the condenser of the recirculation air is bypassed, which is realized using two magnetic valves. In this mode, the superheated vapor directly flows into the condenser of the supply air. In both modes, subcooling is realized in the condenser of the supply air, and thus, a refrigerant receiver is not necessary. Defrosting is achieved using a hot-gas bypass that is not shown in the figure. An enthalpy membrane exchanger is used in the ventilation unit (see Fig. 3). The volume flow of the supply air can be controlled between 60 and 120 m3/h. recirculation air can be controlled between 0 and 135 m3/h. The outdoor unit (see Fig. 3) includes the evaporator, compressor,

Qcond, sup + Qcond, rec + Qloss, comp + Qloss, pipes = Qevap + Pel, comp COPHP, sys =

COPsys =

(1)

Qcond, sup + Qcond, rec (2)

Pel, comp + Pel, inv + Pel, vent , amb + Pel, vent , rec Qcond, sup + Qcond, rec + Q vent

(3)

Pel, comp + Pel, inv + Pel, vent , amb + Pel, vent , rec + Pel, vent

Qcond, rec = (hrec , out

hrec, in )· msup

Pel, vent , rec = (href , HG2

href , TP )· mref (4)

Qcond, sup = (hsup

hsup0 )· msup

Fig. 2. Scheme of the system showing the states of air and refrigerant. 3

Pel, vent , sup = (href , TP

href , sub)· mref

(5)

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Fig. 3. Scheme of the (top view) indoor and (front view) outdoor units.

Qevap = (hamb2· mamb2 + hexh · mexh + Qloss, comp + Qloss, inv + Pel, vent , amb) hexh, h· mexh, h = (href , suc

href , HG = href , HG *

(6)

href , evap)· mref

= (href , HG2

Qvent = (hsup0

HG1

log

(

HG

HG1

amb

HG

amb

)

+ UApipe, HG, room ·

HG 2

log

(

HG1

HG 2

room

HG1

room

)

Pel, inv = Qloss, inv = Pel, comp·(1/

href , HG )· mref

(7)

hamb)·msup

(8)

3.3.1. Overall balance For the evaporator and condensers, PFTHEs with plane fins are used. The PFTHE is discretized using 2·n elements (with n elements used for air and refrigerant). Each element contains fins. The bends are not separately considered because they are not in contact with air. The fin efficiency and flow direction are considered in the calculation of the heat-transfer coefficient of air. Usually, PFTHEs have parallel refrigerant loops (three in this study), which are considered in the model by dividing the heat exchanger into subsections. Fig. 5 shows visualization of the energy balance of one element. Eqs. (15) and (16) can be used for the energy balance of the air and refrigerant flows, respectively. The term with the condensate appears of course only in the evaporator. Eq. (17) expresses the total heat transferred.

Mass flow of the refrigerant mref is calculated depending on speed of the compressor N using Eq. (9).

N ·Dc · 60

(9)

ref , suc

Electrical power consumption of the compressor Pel, comp is calculated according to Eq. (10).

Pel, comp = (href , HG *

(10)

href , suc )· mref

The theoretical outlet temperature of the hot gas is calculated using Eq. (11).

href , HG * = href , suc + 1/

ise ·(h ref , HG, ise

href , suc )

(11)

ref , H

amb)

hair , j · mair = hair , j + 1· mair + UT , j Aref , j (

ref , j

air , j)

href , j ·mref = href , j + 1· mref

ref , j

air , j)

UT , j Aref , j (

mcondensate · h w

(15) (16)

n

Qcond (evap) =

To account for heat losses of the compressor, the enthalpy of the hot gas is calculated using Eqs. (12) and (13).

Qloss, comp = UAcomp ·(

(14)

1)

3.3. Physical model of the PFTHE

3.2. Physical model of the compressor and inverter

vol ·

inv

The volumetric efficiency, isentropic efficiency, thermal-loss coefficient UA comp , and inverter efficiency are measured. The characteristics of the compressors used in this study have been determined by a separate test rig. The corresponding measurement results can be found in the dataset in [16]. Volumetric efficiency vol and isentropic efficiency ise are considered as third-order polynomials, which were obtained from [dataset] [16]. For each speed of the compressor N , one polynomial, which depends on pressure ratio = pHP /pLP and condenser temperature is established.

Fig. 4 shows the calculation flowchart. To account for the ventilation system and the enthalpy recovery, the model in [15] is used to calculate hsup0 and hexh . Low pressure pLP is adjusted until desired superheating temperature TSH is reached. Based on this calculated refrigerant charge, high pressure pHP in the model is adjusted until refrigerant charge mref , set is attained. Refrigerant charge mref is calculated according to [16]. Because of the absence of an accumulator, the subcooling temperature depends on the compressor speed and pressure ratio. The simulations are carried out using MATLAB and Optimization Toolbox. Function fminsearch and the user-defined abort criteria are used. The overall error of the simulation model is approximately 0.15 K for the air temperature and 100 J/kg for the refrigerant enthalpy.

mref =

(13)

mref

To account for inverter efficiency inv , Eq. (14) is used to calculate electrical power of the inverter Pel, inv and the heat losses of the inverter Qloss, inv .

Qloss, pipes, HG = UApipe, HG, amb ·

Qloss, comp

UT , j Aref , j ( j =1

ref , j

air , j)

(17)

To determine the temperature of the refrigerant in Eq. (16), the pressure and its aggregate state (liquid, two-phase, gas) must be known.

(12) 4

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To calculate the air and refrigerant properties, the CoolProp library [17] is used. 3.3.2. Overall heat transfer Overall heat transfer coefficient UT , which corresponds to Aref , can be calculated using Eq. (18).

1

UT =

Aref Afin

1

·

dtube

+

air

tube

1

+

(18)

ref

(19)

Aref = Do · ·Ltube Afin = 2· nfin· Hfin· Lfin

ntube ·Do2 /4·

(20)

3.3.3. Heat-transfer coefficient of air The method defined by [18] is used to calculate the heat-transfer coefficient of air air . In the present study, the correlations in [19] fit well the dimensions of the used PFTHE. A more detailed explanation on how the geometry has to be considered can be found in [19]. A homogenous parallel flow to the fins is assumed. A recent review on calculating compact heat exchangers has been done by [20], and different correlations for different application cases and types of compact heat exchangers can be found there. Heat-transfer coefficient air can be calculated based on the Nusselt number, as expressed in Eqs. (21) and (22).

Nu dh

(21)

Nu = jH · Re· Pr 1/3

(22)

air

=

In this case, L and dh are equal, and therefore, outer diameter Do of the tube must be used. The value of Colburn factor jH must be determined through experiments or validated computational fluid dynamics calculations. In this study, the experimental results from [19] are used. The calculation of jH for a PFTHE with three tube rows is realized using Eq. (23). 0.369·(S / S )0.106 ·(s / D )0.0138 ·(S / D )0.13 t l o t o

jH = 0.163· Re

(23)

3.3.4. Heat-transfer coefficient of the refrigerant To describe the heat transfer of the gas phase, i.e., ref , G or liquid phase, i.e., ref , L , the Dittus–Boelter equation [Eq. (24)], is used. ref , G

=

ref , L

= 0.023·Re 0.8 ·Pr 0.4· / dh

(24)

Because L and dh are equal, inner diameter of the tube Di must be used. In the two-phase region, the correlations in [21] are used to calculate heat-transfer coefficient ref , TP for boiling. The flow direction of the refrigerant is horizontal.

Fig. 4. Flowchart of the calculation of the refrigerant cycle.

ref , TP

=

=

·

(25)

ref , L

1.8 Co0.8

Co =

1 x

(26)

1

0.8

·

g

0.5

(27)

l

In the two-phase region, the correlations in [22] are used to calculate heat-transfer coefficient ref , TP for condensation. ref , TP

=

ref , L·

(1

x )0.8 +

3.8·x 0.76 ·(1 x )0.04 pr0.38

(28)

More complex correlations in the two-phase flow heat-transfer coefficients can be found in the literature. However, because of the high ratio between the heat-transfer coefficient of air and the refrigerant, higher accuracy is not necessary.

Fig. 5. Energy balance of discrete element j .

5

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3.3.5. Pressure drops To account for the pressure drops, a simplified homogeneous-flow model is used. If the liquid and gas phases move with the same velocity, the two phases can be treated as a pseudo single-phase fluid, and Eq. (29) can be used [23]. The geodetic height differences are neglected.

pref , jj + 1 = pref , jj

0.158· Re

0.25·

G2 · dh

Table 1 Measurement points.

(29)

3.4. Measurement

Characteristic

Value

Volume flow supply/extract air Volume flow recirculation air Volume flow additional ambient air Ambient air (temperature and relative humidity) Extract air (temperature and relative humidity) Compressors

80 and 100 m3/h 0, 65, 100, and 135 m3/h 35, 100, and 200 m3/h −10 to +12 °C, 60% to 80%

Compressor speed

Fig. 6 shows the schematic of the used test rig with the proposed heat pump, including the ventilation system. The air ducts are connected to two climate chambers. One climate chamber simulates room air, and the other simulates ambient air. Both climate chambers can be completely controlled relative to the temperature (heating and cooling) and moisture (humidification and de-humidification). The test device is installed in a third air-conditioned room. For all air streams, the temperature is measured near the device using five temperature sensors ( ). In addition, the relative humidity ( ) and temperature are measured after a short distance to ensure sufficient mixture of air. An external pressure drop ( p ) can be applied using an orifice. Additional auxiliary ventilators (not shown in the figure) ensure a constant pressure difference ( p0 ) against ambient in front of the orifice. The volume flows (V ) are measured using the orifice plates and differential pressure sensors. Depending on the range of the expected volume flow, different orifice plates are used. The electrical power consumptions of the fans in the ventilation part, compressor (including the inverter), and additional fans for the ambient air are

21 °C, 50% Hitachi HIGHLY WHP3240BSK and WHP01900BCK 1500–6500 rpm and 1500–7000 rpm

recorded. To obtain additional information about the refrigerant cycle, absolute-pressure sensors are installed before and after the compressor. Moreover, the surface temperatures of the refrigerant pipes are measured using thermocouples installed at relevant positions (compare Fig. 2). All air ducts used to connect the heat pump to the test rig are insulated using a 13-cm-thick layer of insulation. The indoor unit consists of polystyrene with a thickness of at least 25 mm. The outdoor unit is insulated with at least 27 mm of Armaflex. Table 4 in Appendix B lists the measurement devices used and their measurement accuracy. The measurement data, which consist of 446 data points (compare Table 1), can be found in [dataset][24]. Each data point represents a steady-state result, which is the mean value over a 20-min period. The

Fig. 6. Schematic of the experimental setup. 6

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heat pump operates under the desired conditions for at least another 20 min before the measurement is started. Within this period, the standard deviation of the heating power must be below 5%. The time step of the data acquisition is 250 ms, and the mean value is recorded every 5 s. Calculation of the thermal power values is carried out using the equations presented in Section 3.1. The thermodynamic properties are determined using the CoolProp library [17]. 4. Results and discussion 4.1. Performance analysis Fig. 7 shows the steady-state results of the four characteristic system states for each system with 2.5 kW heating power. The heating power and COPHP, sys are plotted at an ambient air temperature from −7 °C to +10 °C for different compressor speeds (starting from 1500 rpm). The results are generated from the measurement results by fitting a polynomial surface using a linear least square algorithm. The expected model error for the heating power is ± 61.1 W , and that for COPHP, sys is ± 0.11 at a confidence interval of 95%. All results exclude the de-icing. By applying a building and system simulation, the additional de-icing power consumption and resulting total heating power can be estimated, similar to that in [9]. Fig. 7(a) shows the first mode of operation for a 100 m3/h supply-air flow only. The recirculation volume flow is set to zero. A minimum additional ambient air flow of 35 m3/h for the outdoor unit and exhaust-air flow of the ventilation system are used as the source. The maximum heating power is limited to approximately 1.2 kW, and COPHP, sys = 1.6 can be expected. The power and performance are quasiindependent of the ambient air temperature. However, even at the lowest compressor speed, COPHP, sys is never higher than 3.0. Fig. 7(b) shows the second mode of operation with an additional recirculation air flow of 100 m3/h. The heating power can be increased to approximately 2.0 kW with COPHP, sys of approximately 2.0. Between 5500 and 6500 rpm, the heating load does not further increase because of the

Fig. 8. Heating power and system COP, including the ventilation system, for the 2.5-kW heat pump under different compressors speeds at ambient-air temperature, 100 m3/h supply, and recirculation air with 200 m3/h additional ambient air.

power limitation of the source. Fig. 7(c) shows the results with 200 m3/ h additional ambient air. The heating power increases to more than 2.5 kW with COPHP, sys = 1.7. The system performance is almost not influenced by the ambient air temperature. Comparison of operating modes (b) and (c) at low compressor speeds of up to approximately 4000 rpm reveals that it is more efficient not to use additional ambient air, which simultaneously allows for the same heating power and higher performance. Above 4000 rpm, additional ambient air is necessary to increase the heating power. Fig. 7(d) shows the same mode of operation as that in (b) but without the exhaust air and using only 300 m3/h ambient air as source. Compared with (c) not only does the heating power decrease by 20% at higher compressor speeds but the COPHP, sys also decreases by 15% at low compressor speeds. The lower source temperature even leads to an increased volume flow at a lower evaporating temperature and thus

Fig. 7. Heating power and system COP for the 2.5-kW heat pump under different compressor speeds as a function of the ambient air temperature. (a) 100 m3/h supply air with no additional ambient air. (b) 100 m3/h supply and recirculation air with no additional ambient air. (c) Similar to (b) with 200 m3/h additional ambient air. (d) 100 m3/h supply and recirculation air with no exhaust air but only 300 m3/h ambient air. 7

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Fig. 9. Comparison of the measured and simulated results of the 1.4-kW heat pump with 100 m3/h supply, recirculation air, and 100 m3/h additional ambient air with and without thermal losses. (a) Total heat power. (b) System COP of the heat pump.

Fig. 10. Optimization study of the 1.4-kW heat pump with 100 m3/h supply, recirculation air, and 100 m3/h additional ambient air (

amb

= 0 °C , N = 7000 rpm ).

recirculation air, and 100 m3/h additional ambient air. The results with and without thermal losses in the refrigerant cycle are presented. The thermal-loss coefficients of the refrigerant cycle [compare Eqs. (1) and (7)] are determined from the measurement results using linear regression and are listed in Table 3, Appendix A. The dotted lines show the expected measurement uncertainty (approximately 8%). A good agreement can be observed for all results, and most of the data points are within the confidence interval. Without considering the thermal losses, all results would be slightly outside the confidence interval. Various reasons can be cited for these deviations. For instance, the air distribution through the evaporator is not optimized, which possibly leads to smaller efficiency [25]. This limitation can be improved in several ways, e.g., by adding an air distributor to the outdoor unit, similar to that in [26] or [27].

reduced heating power. Moreover, the ambient-air ventilator consumes more power because of the higher necessary volume flow; therefore, the COP further decreases. For the heat pump with 1.4-kW heating power, the same trends can be observed. However, the minimal provided heating load decreases from 800 to 470 W at an ambient air temperature of +10 °C. Simultaneously, COPHP, sys increases from 4.0 to 4.9. The maximum heating load decrease from 2800 to 1430 W at an ambient air temperature of 0 °C with COPHP, sys of approximately 2.7. Fig. 8 shows the overall thermal power, including the ventilation, for system configuration (b) shown in Fig. 7. The minimal thermal power of the system is approximately 1230 W with COPsys = 4.6. The thermal power at an ambient-air temperature of 0 °C and maximum compressor speed reaches 3540 W with COPsys = 2.1. Owing to the enthalpy exchanger, COPsys increases at lower ambient air temperatures. Thus, for an ambient air temperature of −7 °C, COPsys at the maximum compressor speed of 6500 rpm increases from 2.1 to 2.5. At the same time, the thermal power decreases from 3540 to 3200 W.

4.3. Optimization by simulation The validated model is used to demonstrate the optimization potential of the system. Fig. 10 shows the results of the heating load of the supply, recirculation air, and COPHP, sys at a compressor speed of 7000 rpm, an ambient temperature of 0 °C, and an initial refrigerant charge of 450 g. The initial refrigerant charge is defined as the instance when only liquid refrigerant enters the expansion valve in any relevant mode of operation at the lowest compressor speed. A critical factor in the design of the refrigerant cycle of a speed-

4.2. Comparison with the simulation results The simulation model was validated using the measurement data points. Fig. 9 shows an example of the comparison of the heating power and COPHP, sys of the 1.4-kW heat pump with a 100 m3/h supply, 8

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controlled air-to-air heat pump is the refrigerant charge. Varying it by simulation shows that the optimum value has already been reached. Subcooling temperature of 16.5 K can be achieved. An increase in the refrigerant charge (and an increase in the subcooling temperature to 19 K) leads to a change in efficiency of less than 1%. The total heating power slightly increases by 15 W with an additional refrigerant charge of 75 g (subcooling temperature of 22.5 K). However, the heating power of the supply-air condenser decreases by 7% because a considerable part (almost 45%) of the condenser is filled with liquid refrigerant, which leads to a reduced heat-exchanger efficiency. A 25-g lower refrigerant charge leads to slightly lower efficiency and total heating power. However, the refrigerant cannot be guaranteed any more to be in the liquid-phase at the inlet of the expansion valve under different boundary conditions. The relatively high refrigerant charge without an accumulator leads to a reduced heat-exchanger efficiency of the supply-air condenser. The optimization shows that an increase in the heat-exchanger area (air and refrigerant sides in both condensers) can improve the efficiency by 5%, see Fig. 10. In this optimization, the subcooling temperature is kept at 16.5 K. The decrease in the temperature difference and the pressure ratio account for the highest effect on these results. In the presented configuration the heat losses of compressor and inverter can be used by the evaporator (compare Fig. 2). However, if these losses would not be used at all, the heating power would decrease by 53 W and the COP by 1.9%. From the thermodynamic point of view compressor and inverter could also be situated in the supply air flow after the condenser. This would increase the total heating power by 110 W and the COP by 11.5%. However, this is not the preferred solution, as it is difficult to reduce the noise emissions of the compressor.

5. Conclusions A novel exhaust-air to supply-air heat pump combined with a ventilation system and recirculation air was presented. Two functional models with a heating power of 1.4 kW and 2.5 kW, respectively, were built and tested in the laboratory. The measurement results consisting of 446 measurement points were used to validate a simplified physical model. The heat losses in all components were considered in detail using calibration, which had high relevance because of the low overall heating power of such systems. Using the validated model, optimization potential of approximately 5% was achieved by increasing the condenser size. This model can be used to support further optimization. In particular, the system needs to be further developed to use it for cooling purposes. The overall system efficiency under dynamic conditions and considering de-icing will be shown and compared with other systems by means of building and system simulations. Acknowledgements This work is part of the Austrian research project FiTNeS “Fassadenintegrierte modulare Split-Wärmepumpe für Neubau und Sanierung” (2018-21); Förderprogramm Stadt der Zukunft 5. Ausschreibung 2017, FFG Austria, Project number: 867327. Declaration of Competing Interest None.

Appendix A See Tables 2 and 3. Table 2 Components. Characteristic

Value

Enthalpy Exchanger Manufacturer Product Height Temperature Effectiveness Moisture Effectiveness

CORE Energy Recovery Solutions C-HRV366 185 mm 82% (21 °C/4 °C, 50%/80%) 60% (21 °C/4 °C, 50%/80%)

Ventilator indoor unit Manufacturer Product

ebmpapst RadiCal R3G133-RA01

Ventilators outdoor unit Manufacturer Product

ebmpapst RadiCal R3G133-RA01

Compressors Manufacturer Series Product Displacement Refrigerant Fill charge Insulation

Hitachi HIGHLY WHP3240BSK 15.1 cm3/rev R134a 450 g ± 20 g 28 mm Armaflex

Inverter Manufacturer Product

LSCONTROL SpeedControl 1045E PMSM097 1.5 kW

Electronic Expansion Valve Manufacturer Product Superheating

CAREL CAREL E2V at least 5 K

WHP01900BCK 10.2 cm3/rev R134a 450 g ± 20 g

Condenser (supply and recirculation)

(continued on next page) 9

Applied Thermal Engineering 162 (2019) 114230

D. Siegele, et al.

Table 2 (continued) Characteristic

Value

Manufacturer Type Rows Width/Height Hfin /Length Lfin Tube diameter Di /thickness dtube /material Length tubeLtube Area fins Fins thickness t fin /material Fins numbers nfin /space sfin

not specified PFTHE 3 205/180/65 mm

6 mm/0.28 mm/Cu

12.0 m 2.0 m2 0.10 mm/Al

85/2.28 mm

22.5/22.5 mm

Pipes distance St /Sl

Evaporator Manufacturer Type Rows Width/Height Hfin /Length Lfin Tube diameter Di /thickness dtube /material Length tube Area fins Fins thickness t fin /material Fins numbers nfin /space sfin

not specified PFTHE 3 200/500/130 mm

7 mm/0.28 mm/Cu 36.0 m 8.2 m2 0.18 mm/Al

63/2.95 mm

22.5/22.5 mm

Pipes distance St /Sl

Table 3 Refrigerant pipes and built-in parts (approximate lengths). From

To

Diameter

Length

Boundary condition

Insulation Thickness

UA

Remark

Compressor

Magnetic valves

170 cm 200 cm

Condenser recirculation condenser supply Condenser supply Filter drier

8 mm 8 mm 8 mm 6 mm 6 mm

90 cm 30 cm 100 cm 220 cm 130 cm

6 mm 8 mm 8 mm

90 cm 80 cm 160 cm

ambient room room room room room room ambient ambient ambient ambient ambient ambient ambient

8 mm 8 mm n/a 6 mm 6 mm 6 mm 6 mm 6 mm 6 mm n/a 6 mm 6 mm 6 mm 28 mm

0.46 W/K 0.30 W/K – – – – 0.67 W/K 0.44 W/K – – – – – –

additional thermal bridges at connections

Magnetic valves Magnetic valves Magnetic valves Condenser recirculation Condenser supply

8 mm 8 mm

Filter drier Sight glass Sight glass EEV evaporator compressor

Volume 0.122 l not further specified EEV evaporator compressor

additional thermal bridge

additional thermal bridge additional thermal bridge only partly insulated

Appendix B See Table 4. Table 4 Measurement devices and accuracy. Measurement value

Pos.

SI unit

Measurement device

Measurement accuracy

Temperature

°C

Pt100 4-wire with NI 9217 Thermocouples type K with NI 9213

Pt100: ± 0, 05°C TC: ± 0, 10°C

Humidity ratio

x

% g/kg

Enthalpy

h

kJ/kg

Relative humidity

Volume flow

Mass flow Absolute pressure Differential pressure Electrical power consumption Thermal power

°C

V m pabs p, Pel Q

m3/h

p0

kg/s Pa Pa W W

E + E EE210 Pt100 and capacitive humidity sensors

: ± 0, 10 °C : ±1, 5%

Calculated

@ 21° C , 50% : ± 0, 24 g/kg

Calculated

@ 21

FläktWoods MR Huba Control 699

@ 4 °C , 60% : ± 0.36 kJ/kg ± 5.0%

@0

Calculated THIES Clima Barogeber Huba Control 699 Iskra MT 400 Calculated

10

± ± ± ± ±

°C ,

60% : ± 0, 06 g/kg

°C ,

5.0% 25 Pa 2 Pa 0.5% 8.0%

50% : ± 0, 87 kJ/kg

Applied Thermal Engineering 162 (2019) 114230

D. Siegele, et al.

Appendix C. Supplementary material Supplementary data to this article can be found online at https://doi.org/10.1016/j.applthermaleng.2019.114230.

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