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Sizing strategy of on–off and modulating heat pump systems based on annual energy analysis
international journal of refrigeration 65 (2016) 183–193
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Sizing strategy of on–off and modulating heat pump systems based on annual energy analysis G. Bagarella *, R. Lazzarin, M. Noro Department of Management and Engineering, University of Padua, Str. lla San Nicola 3, Vicenza, 36100, Italy
A R T I C L E
I N F O
A B S T R A C T
Article history:
In recent years the use of high efficient variable speed heat pumps has spread widely. Though
Received 15 October 2015
several studies were carried out in the past comparing on–off and modulating strategies
Received in revised form 16
in heat pump units, only few works considered the whole heat pump system. Moreover,
February 2016
because of lack of knowledge about cycling losses, most of comparisons between modu-
Accepted 16 February 2016
lating and on–off strategies did not consider the effects of the reduced coefficient of
Available online 23 February 2016
performance of a system at start-up. In this work simulations were carried out considering both on–off and modulating air-to-water heat pump systems in a single family house
Keywords:
located in Italy. The aim of this study is to evaluate how the heat pump sizing, the thermal
Heat pump systems
storage sizing and the cycling losses phenomenon can influence the annual energy perfor-
On–off
mance of both the systems (on–off and modulating) and to investigate when modulating
Modulation
heat pump plants are actually more efficient than on–off.
Variable speed
© 2016 Elsevier Ltd and IIR. All rights reserved.
Sizing strategy
Stratégie de dimensionnement de systèmes de pompe à chaleur tout ou rien et à modulation basée sur une analyse énergétique annuelle Mots clés : Systèmes de pompe à chaleur ; Tout ou rien ; Modulation ; Vitesse variable ; Stratégie de dimensionnement
1.
Introduction
In recent years the use of high efficient modulating heat pumps (equipped with inverter driven variable speed compressors) has spread widely. From the theoretical point of view several issues
* Corresponding author. Tel.: +39 3478037687; Fax: +39 0444 998888. E-mail address: [email protected] (G. Bagarella). http://dx.doi.org/10.1016/j.ijrefrig.2016.02.015 0140-7007/© 2016 Elsevier Ltd and IIR. All rights reserved.
should be considered when a fixed speed heat pump (here referred to also as on–off heat pump) is compared to a variable speed one (referred to also as capacity controlled or modulating heat pump). Peak load in building heating usually takes place only at the outdoor design temperature (e.g. −5 °C), that is for some tens
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Nomenclature BLDC COP CL Kp PID SPF Tc Td Te Text Ti tm,cycle tm,on
tm,off
two TEV TL
brushless direct current motor coefficient of performance cycling losses [kJ] proportional gain of PID control proportional, integral, derivative control seasonal performance factor condensation temperature [°C] derivative time of PID control [s] evaporation temperature [°C] temperature of the external air [°C] integral time of PID control [s] average water temperature at the condenser outlet during the on–off cycle [°C] average water temperature at the condenser outlet during the on period of an on–off unit [°C] average water temperature at the condenser outlet during the off period of an on–off unit [°C] water temperature at the condenser outlet [°C] thermostatic expansion valve tank thermal losses [kJ]
of hours per year. If the selection of heat pump capacity is based on peak load, for most of the time heat pump capacity shall be controlled as it exceeds heating requirements; a fixed speed heat pump controls capacity by switching on–off. Undersizing an on–off heat pump with respect to the peak load lowers control requirements, however without eliminating them. Of course, undersizing heat pump capacity increases the number of operation hours of the auxiliary system (Lazzarin, 2012). As mentioned above, capacity control of a fixed speed heat pump gives rise to a sequence of on–off cycles that lower the efficiency of the unit. A variable speed unit, on the other hand, could run continually, so reducing the on–off frequency and the cycling losses phenomenon (Bagarella et al., 2013).
In principle, a modulating heat pump is more efficient than an on–off unit of the same nominal heating capacity. In fact, the same heating energy during part load operation is given by the modulating heat pump and by the on–off one, but the latter works only for a fraction of time, whereas the former operates continuously. Hence, the average temperature of the secondary fluid, during the on period, must be higher for the on–off unit to allow the due heat exchange. Fig. 1 depicts the typical water temperature profile at the outlet of an on– off heat pump unit. When the heat pump is activated (on period), the water temperature increases. As soon as the water temperature reaches the upper limit (whose value depends on the control settings) the unit is switched-off by the controller. Then, the water temperature decreases, due to the building thermal load. Finally, once the temperature reaches the lower limit (which also depends on the control settings) the heat pump unit is switched on again. In Fig. 1, the outlet water temperature of a modulating unit can be evaluated as the average water temperature of the on–off one (tm,cycle) during the whole period, which is clearly lower than the average water temperature of the on–off unit during the on period (tm,on). This will affect the condensation temperature the same way, leading the modulating heat pump to operate with a lower condensation temperature, then with higher COP (coefficient of performance, ratio of heating provided to work required). From the previous rationale a meaningful increase in thermodynamic efficiency is waited when considering a variable capacity heat pump towards a fixed one. In fact, Aprea et al. (2006) compared an on–off air-water heat pump with a modulating unit and found a 20% electric energy reduction for the latter. Energy savings were mainly attributed to the reduction of the compression ratio during the modulation phase (Aprea et al., 2006). However, a thorough energy analysis of the system shall consider the electricity needs of the inverter that equips a variable speed heat pump. Though several studies (e.g. Aprea et al., 2006, Shao et al., 2004, Zhao et al., 2003) have been carried out in the past comparing on–off and modulating strategies in single heat pump units, only few studies (Adhikari et al., 2012; Cheung and Braun, 2014; Lee, 2010; Madani et al., 2011b) analyzed the whole heat pump system, which includes not only the single heat pump
Fig. 1 – Outlet water temperature profiles in on–off heat pump unit (Fahlen and Karksson, 2005).
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unit, but also the heat source, the heat sink, the heating distribution system, the thermal storage, the liquid pump and the auxiliary heaters, as explained in section 2. Lee (2010), Adhikari et al. (2012) and Cheung and Braun (2014) simulated annual performance of both on–off and capacity controlled heat pump systems in different climates and concluded that a capacity controlled system can achieve considerable energy savings compared to an on–off one, especially in mild climates. Anyway, these studies did not consider, for a single climate, the possibility of sizing the on–off system to satisfy different fractions of the building peak load. Does the energy saving of a modulating heat pump system (for a given climate) depend on the size of the benchmark on–off system? Only Madani et al. (2011b) raised this issue, simulating the annual operation of both on–off and modulating geothermal heat pumps with the typical thermal load of a single house located in Stockholm (Sweden). Assuming as auxiliary system an electric heater, they concluded that the modulating heat pump led to a better performance only when the on–off heat pump was sized to cover less than 65% of the heating peak demand. On the other hand, when the on–off heat pump presented a higher capacity, no energy savings were found because benefits deriving from a lower mean compression ratio allowed by the modulating heat pump were completely compensated by the energy requests of both inverter and pumps. In this paper, as in Madani et al. (2011b), several simulations were carried out considering on–off and modulating heat pump systems (with different capacities) with the typical thermal loads of a single family house located in Padua (NorthEast of Italy). The main difference between this study and the above cited (Madani et al., 2011b) is in the heat pump source, since outside air (instead of ground) is here considered. Heat pump capacity is strongly influenced by the heat source temperature and the annual variations of the air temperature is much higher than the maximum variation of the ground temperature (e.g. 3 K was the maximum annual ground temperature variation considered by Madani et al., 2011a). Moreover, when the air temperature decreases with a reduction of heat pump capacity, the building thermal load usually increases. There-
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fore, the heat-pump capacity modulation ability assumes a crucial importance in air-to-water applications, whereas it represented a minor problem in the ground source system studied by Madani et al. (2011a, 2011b). Another major difference is in the choice of the auxiliary heater for the heat pump system. Whereas in Sweden electric heaters are common, probably due to favorable tariffs, gas boilers are used in Italy. Therefore, a condensing boiler (fueled by natural gas) was considered. Finally, probably because of lack of knowledge about the cycling losses phenomenon, most of the comparisons between modulating and on–off strategies found in literature (including the study proposed by Madani et al., 2011b) do not consider the effects of the reduced COP of a system at start-up. In this paper cycling losses penalizations have been duly considered basing on a previous work specifically devoted to this computation (Bagarella et al., 2013).
2.
Methodology
To simulate the annual behavior of both on–off and variable capacity heat pumps the system was firstly modeled. Fig. 2 shows a schematic overview of the modeled heat pump system. The air-to-water heat pump unit supplies heat to the building by means of a hot water storage tank and is modeled by linking together the sub-models of condenser, expansion valve, evaporator and compressor as in Madani et al. (2011b). Some technical data of the considered heat pump unit are summarized in Table 1. REFPROP 9.0 (Lemmon et al., 2013) was used to estimate the refrigerant properties in each operating condition. Heat transfer coefficients of both evaporator and condenser were set according to results obtained by tests carried out with an equivalent experimental unit as in Madani et al. (2011c). Superheating and subcooling were considered constant as in Cecchinato et al. (2010). The global electric consumption of the system was calculated summing up compressor, inverter, pump and fan require-
Fig. 2 – Schematic overview of the heat pump system.
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Table 1 – Heat pump unit data. Component Condenser
Evaporator
Refrigerant Nominal heating capacity (Text = −5 °C) Expansion valve Compressor
Table 2 – Building data. Information
Plate heat exchanger Heat exchange surface: 0.79 m2 Plates: 20 Finned coil Wavy fins Tube length: 920 mm Number of columns: 48 Number of rows: 2 Fin thickness: 0.15 mm Fin pitch: 1.95 mm Tube spacing: 25.4 mm R410A 5.9 kW TEV Kv = 0.066 m3 h−1 BLDC Twin-rotary Displacement: 1.3 dm3
ments. The compressor consumption was estimated basing on manufacturers’ data for each operating conditions. Pump and fan consumption (0.03 kW and 0.06 kW respectively) was assumed constant. The inverter efficiency was considered in a range between 88% and 95%, depending on the compressor rotation speed, as specified by the manufacturer. Cycling losses were measured by specific experimental tests as described in Bagarella et al. (2013) and results were included in the model. In that study, details about how the number of hourly start-ups affects the percentage efficiency loss of the unit are reported and deeply discussed. One of the main conclusions of that study is that the cycling losses phenomenon strongly depends on the expansion valve used. If a thermostatic expansion valve is used, the COP of the unit at the start-up is strongly reduced. This behavior is not due to a higher electric power absorbed by the compressor, but rather to the poor heating effect during the starting transient. Experimental tests demonstrated that, once the unit is switched on, steady state conditions are attained only after 100–120 seconds. During this transient period, the average COP is approximately 50% lower than the steady state value. Basing on this consideration, the algorithm proposed by Madani et al. (2011b) was modified, and the unit was forced to operate with a reduced COP (50% lower than the one calculated through Madani’s model) during each start-up period (110 seconds). The compressor manufacturer claims a maximum number of hourly on–off cycles up to 12. Nonetheless, the maximum number of startups here considered is 5, as more cycles would excessively reduce the compressor life. A simplified method to account for frost formation and defrost cycle consumption was also implemented in the model. Details about this simplified model (i.e. information regarding main assumptions, software used to roughly account for frost formation and model validation) are reported in Bagarella et al. (2016). The average fin temperature was estimated using software EVAP-COND developed by NIST (NIST, National Institute of Standard and Technology, 2003). When fin temperature gets conditions that favor frost formation (fin
Parameter Building type Building volume Building height Vertical structures transmittance Horizontal structures transmittance Floor transmittance Windows transmittance Ventilation rate Set point indoor temperature Peak load (Text = −5 °C) Annual thermal energy demand
Information Single family house (2 floors) 363.5 m3 5.5 m 0.28 W m−2 K−1 0.24 W m−2 K−1 0.27 W m−2 K−1 1.60 W m−2 K−1 109 m3 h−1 20 °C 6.1 kW 5757 kWh
temperature below 0 °C and below outdoor air dew point), frost starts building up. Once the frost has reached its upper bound (average frost thickness 0.7 mm) defrost (by electric heaters) is activated. The model of the heat pump unit was validated comparing its behavior against those of the real modulating one. In particular, a four hour test in a climatic chamber was carried out (main characteristics of the real unit are described in Table 1) according to test conditions suggested by Ente Nazionale Italiano di Normazione (2012) (National Standard Agency) for an average European climate. During the test, for each partload condition suggested by Ente Nazionale Italiano di Normazione (2012), the following quantities were measured and compared with those calculated by the model (with the unit in steady-state condition): evaporation and condensation temperatures, compressor frequency and thermal power. At the end of the test, the electric energy absorbed by the capacity controlled unit (5.9 kWh) was compared to the electric energy calculated by the dynamic model (5.7 kWh). The same procedure was followed to validate the on–off unit (the compressor of the real capacity controlled unit was forced to operate at its maximum speed during the on periods, regulating its capacity by switching the compressor on–off to produce the same thermal energy as the capacity controlled unit). The electric energy absorbed by the real on–off unit (7.5 kWh) was 5% higher than the electric energy calculated by the dynamic model (7.1 kWh). The building (Table 2) was modeled in TRNSYS 16.0, estimating the thermal loads under annual climatic data (Test Reference Year, TRY, 1986) of Padua (Italy). The design temperature is −5 °C and heating degree days are 2381. The thermal properties of the building envelope are those compliant to the present National Regulations for new buildings’ standards. The thermal load of the building and the thermal capacity of the heat pump unit were used to continuously calculate the temperature of the hot water tank (Fig. 2). The fully-mixed thermal storage decouples the heat produced by the heat pump from the heat required by the building. The storage protects the heat pump unit from an excessive number of on–off cycles and ensures the thermal comfort inside the building avoiding excessive variations in the air temperature during defrosts. A low temperature radiant heating distribution system was considered. The set-point of the hot tank was set at 35 °C.
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Fig. 3 – Compressor isentropic efficiency at different frequencies and evaporation (Te) and condensation (Tc) temperatures.
The control unit allows the heat pump to work either with on–off cycles or with a capacity control algorithm. The control variable was the temperature of the thermal storage. In the case of on–off strategy a 2 °C dead band around the set-point of the water tank was used, while in the capacity controlled unit a PID algorithm was used to set the compressor speed. When the heating capacity of the heat pump unit fails to cover the heating demand, a condensing boiler auxiliary heater is activated. The condensing boiler (fueled by natural gas) is modeled through its thermal efficiency, which is assumed a function of the water temperature at the boiler inlet. The thermal efficiency profile suggested by Lazzarin (2012) was used. As in Table 1, the nominal thermal capacity of the real heat pump is 5.9 kW (when Text = −5 °C). According to this validated model, other heat pumps of different nominal capacity were simulated as explained in section 5.1. Climatic conditions have two effects on the system: lower air temperatures mean lower evaporation temperatures (consequently lower COPs) and, at the same time, higher building heating demand. The global algorithm of the heat pump system was implemented in LABVIEW 8.5 environment by linking together the sub-models of the elements following the approach in Fig. 2.
3. COP variations and modulation capacity of the variable speed unit A first application of the model of both heat pump units (on– off band modulating) and heat pump systems was devoted to study the capacity controlled unit performance. The influence of outside air temperature and heating demand on the heat pump COP was investigated. For a better understanding of the results, compressor isentropic efficiencies (calculated basing on manufacturer data) at different frequencies are reported in Fig. 3. The maximum isentropic efficiency of the compressor is obtained for frequency from 50 Hz to 70 Hz. Similar results were found by Madani et al. (2010). As waited, the isentropic efficiency lowers for higher pressures ratios. However, for very low condensation temperatures (e.g. Tc = 28 °C), that is for a low pressure ratio, the isentropic efficiency lowers as well, probably because the used twinrotary compressor was not optimized for really low pressures ratios. Fig. 4 represents the COP of the modulating unit as a function of the thermal load of the building for five different external air temperatures. The capacity at the maximum frequency
Fig. 4 – Heat pump unit COPs at different temperatures of the external air (Text, which is equal to the evaporator air inlet temperature) and for different thermal loads.
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Fig. 5 – Behavior of both high reactive and high stable compressor PID controls.
(120 Hz) lowers for decreasing outside air temperature. The maximum COP point shifts toward lower capacities too because the compressor is optimized for a given frequency (e.g. 50 Hz), and the lower Text, the lower the thermal capacity at which this frequency is reached. The most important feature of the curves in Fig. 4 is the significant decrease of the COP (after the maximum value) moving toward higher capacities. The modulation ability of an inverter driven heat pump unit should not be overrated. When the nominal frequency of the compressor is 120 Hz, its velocity can be reduced to 20– 30 Hz. In fact, when the compressor frequency is lower than this minimum value, a strong reduction in volumetric efficiency takes place. As a matter of fact, at low rotation speeds the lubrication does not assure a good tightness against the return of refrigerant to the suction. This problem affects all the most widespread small size compressor models. Scroll compressors allow a sensible advantage compared to rotary and twin-rotary compressors, since lower minimum velocities (20 Hz, instead of 30 Hz) are sometimes claimed by manufacturers (Emerson-Copeland, 2015; Siam Compressor Industry, 2010). Moreover, the electric motor efficiency usually reduces for lower frequencies. Furthermore, the higher electric consumption at low rotation speed is converted into thermal power and then transferred to the refrigerant. Hence, the modulation ability is far from being proportional to the reduction in rotational speed of the compressor. This ability is further decreased when the outside temperature is higher, since the heat pump capacity increases just when an effective modulation would be required by the decrease of building thermal load. In addition to the minimum compressor frequency, also the PID parameters of the compressor controller affect the modulating heat pump minimum capacity. As depicted in Fig. 5, the higher the reactivity of the controller (how fast the controller reacts to a variation of the controlled variable, which is the water temperature at the condenser inlet), the higher the overshoot of the compressor frequency (difference between the final and the minimum value of the process variable, which is the compressor frequency), thus the higher the minimum thermal load which can be satisfied by the heat pump unit avoiding the compressor frequency to fall below 30 Hz. In fact, when the heating load decreases, if the reactivity of the controller is too high the
compressor frequency rapidly decreases too and it may fall below the minimum compressor frequency (specified by the manufacturer). In this case, the heat pump is switched off, as a protection of the compressor (especially if this operating condition lasts for several minutes). This usually happens when the proportional gain (Kp) is set to a high value and the integral time (Ti) is set to a low value. On the other hand, the higher the stability of the control, the longer the necessary time to ensure that the heat pump capacity matches the load. Table 3 reports the capacity ratio (the ratio of minimum capacity to the maximum capacity) of the considered compressor (minimum frequency set to 30 Hz) when the parameters of the PID controller are set in order to let the heat pump operate in reasonably stable conditions. The reader should be aware that, while the capacity ratios reported in Table 3 were obtained through a dynamic simulation of the heat pump, the profiles in Fig. 5 were proposed with the only purpose of representing high reactive and high stable controls. To exemplify how the PID parameter setting affects the reactivity of the regulation, coherently with the real settings used in the simulations (Kp = 10, Ti = 180 s; Td = 0 s), the high reactive and the high stable controls in Fig. 5 might be respectively obtained with the following parameters: Kp = 20, Ti = 40 s, Td = 0 s and Kp = 5,Ti = 500 s, Td = 0 s.
4.
Comparison during one whole year
Dynamic simulations were carried out with a 15 minute time step to compare the performance of the plant equipped with
Table 3 – Capacity ratio of the heat pump unit at different temperatures of the external air. Text (°C) 16.0 12.0 7.0 2.0 −7.0
Maximum capacity (kW)
Minimum capacity (kW)
Capacity ratio%
8.2 8.0 7.2 6.7 5.8
4.4 4.0 3.3 2.9 2.2
53.8% 50.0% 46.5% 42.9% 38.2%
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Fig. 6 – Bivalent temperature of the considered heat pump (HP).
on–off or modulating air-to-water heat pump. Firstly a 150 liter thermal storage for both on–off and modulating heat pump was considered. In the first simulation the two heat pump units were sized to cover the 91.4% of the maximum thermal load required for heating up the house. This size of both heat pumps corresponds to a −4 °C bivalent temperature, as reported in Fig. 6. From a theoretical perspective the bivalent temperature may be defined as the minimum temperature at which the thermal capacity of the heat pump matches completely the building thermal load (Ente Nazionale Italiano di Normazione, 2012). When the outside air temperature is lower than the bivalent temperature an auxiliary heater is thus required. On the other hand, when the outside air temperature is higher than (or equal to) the bivalent temperature, the heat pump operates either modulating or in on–off, as the nominal heat pump capacity overrates the load. Fig. 7 shows the different annual profiles of the monthly seasonal performance factor (SPF) of both capacity controlled and on–off heat pump systems. The seasonal performance factor of a heat pump system is defined as the ratio between the seasonal thermal energy delivered by the plant (including auxiliary heaters) and the energy input required by the whole heat pump system (Directive2009/28/EC of the European
Parliament, 2009). For most of the time the heating load is less than the heating capacity of the heat pumps. Under these conditions the capacity controlled heat pump unit can work with higher SPF than the on–off one. The differences in the SPF values of the two systems are higher in spring and autumn, when is higher the need for modulation, as the building load is lower. The capacity controlled unit can work continuously reducing the compressor frequency, thus operating with lower pressure ratios than the on–off unit, consequently with less compressor specific work. However, the differences in the SPF values are lower when the heating loads approach the nominal capacity of the heat pump units. Then the capacity controlled heat pump unit operates at full capacity, without the benefits due to modulation. Fig. 7 also reports the cumulative number of on–off cycles for both heat pump systems. Though the capacity controlled system leads to a drastic reduction of on–off cycles, its modulation capacity decreases when the outdoor air temperature is higher. This phenomenon is confirmed by the profile of the cumulative on–off cycles curves, whose slope increases in spring and autumn with respect to the coldest periods. Fig. 8 illustrates a breakdown of the energy balance of the system. The electric consumption of the variable speed unit
Fig. 7 – Annual profiles of the monthly SPF and cumulative on–off cycles of both capacity controlled and on–off heat pump systems. Monthly thermal energy needs and monthly maximum loads are reported too.
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Fig. 8 – Absolute and percentage breakdown of energy required by the system with both on–off and capacity controlled heat pump systems.
is 1662 kWh, 13.2% less than the electric energy absorbed by the on–off heat pump. The pump and the fan consumption are higher in the capacity controlled system since they operate for a longer time. Moreover, the inverter energy cost must be accounted for in the capacity controlled unit. In the specific conditions of the simulation, the energy cost of defrosting is slightly higher for the capacity controlled unit. This behavior could be attributed to the fact that the modulating unit operates for a longer time during the heating season. Last but not least, the capacity controlled heat pump system strongly limits the total number of on–off cycles, as described in Fig. 7, and so it definitely reduces cycling losses. Fig. 8 suggests that, if annual performances of the two systems are compared, cycling losses cannot be neglected. In the present simulation the effect has been estimated to be 3.4% and 8.3% of the total electric consumption respectively for the modulating and the on–off heat pump systems. These values are higher than the 2.1% efficiency loss reported by Uhlmann and Bertsch (2012), who considered a maximum number of 3 hourly start-ups. The main difference in the two simulations can be identified in the choice of a water tank whose volume is finalized to limit to 5 the maximum number of hourly on–off cycles in this work. In fact, real heat pump systems are usually designed as to avoid the compressor to operate with an on– off frequency too high. The reason is that, even if the effects of the cycling losses phenomenon are often ignored, an on– off frequency too high shortens compressor life. In real
applications, even if the experienced designer has sized the volume of the water storage to avoid a maximum number of hourly on–off cycles higher than 5, a sixth hourly activation may sometimes happen. However, this occurrence is rather rare so that the effects on the compressor life are negligible.
5. Analysis of main variables influencing the seasonal performance 5.1.
The influence of the unit size
To analyze how the sizing of the on–off heat pump can influence system performance, several simulations were run varying the nominal heating capacity of the heat pump, meeting a fraction from 74.3% to 99.8% of the annual heating load. On the contrary, the nominal capacity of the inverter driven heat pump was at first kept constant (5.9 kW). As in Fig. 9, the higher the heating capacity the lower the bivalent temperature (thus the lower the auxiliary energy required). Table 4 reports the annual number of on–off cycles of the on–off heat pump for different fractions of the satisfied demand. The number strongly increases with that fraction and so the cycling losses. Accordingly, an SPF reduction of the unit is observed. On the other hand, the lower the heating capacity of the heat pump considered, the higher the auxiliary energy used.
Fig. 9 – Bivalent temperature depending on the heating capacity of the heat pump unit (HP).
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Table 4 – Total energy, on–off cycles, SPF of the unit, SPF of the system and percentage energy savings of the capacity controlled system for different covering percentages of annual heating demand (or peak load) of the on–off heat pump system. Annual heating demand covering % Peak load covering % On–off cycles Auxiliary energy (kWh) Cycling losses (kWh) Total energy consumption (kWh) SPF on–off HP unit SPF on–off HP system Inverter HP savings
74.3% 33.5% 1237 666 43 1994 3.34 2.89 16.2%
81.7% 38.6% 1725 473 60 1956 3.27 2.95 14.6%
87.% 43.8% 2233 321 78 1931 3.22 2.99 13.4%
91.9% 54.1% 2700 210 95 1914 3.18 3.01 12.7%
94.9% 59.2% 3140 132 110 1909 3.14 3.02 12.4%
97.1% 65.7% 3542 75 124 1906 3.12 3.02 12.3%
98.4% 72.1% 3896 41 136 1909 3.09 3.02 12.5%
99.7% 91.4% 4574 9 160 1924 3.05 3.00 13.1%
99.8% 94.0% 4837 5 169 1932 3.04 2.98 13.5%
Thus the weight of SPF of the auxiliary heater over the SPF of the whole system is growing. As the auxiliary heater performance is always lower than the heat pump, in Table 4 and Fig. 10, the SPF of the on–off heat pump system is always lower than the SPF of the on–off heat pump unit. The highest value of the system SPF was obtained in this study for a fraction of the annual heating demand satisfied from 95% to 98%, which corresponds (in the specific climatic conditions of Padua) respectively to a 59%–72% nominal capacity with respect to the peak thermal load. Fig. 10 and the last row of Table 4 indicate that the energy savings of the capacity controlled heat pump, with respect to the on–off system, always exceeds 12%. Whereas an oversizing of the on–off heat pump, with regard to the above mentioned optimum capacity, slightly penalizes the system, an undersizing is more severe for the low efficiency contribution of the auxiliary. A further study was devoted to determine the best sizing of the inverter driven heat pump. However, every further reduction of its capacity with respect to the above simulation revealed a worse performance. The main reason is that a reduction of cycling losses for lower capacities does not compensate the penalty of the lower performance of the increasing auxiliary operation.
of the heat pump units constant and varying the volume of the tank for both the considered systems. In the simulations the thermal losses of the storage (a 2 °C dead band around the set point was used) have been considered according to Ente Nazionale di Normazione (2008). A minimum 100 liter water storage volume was provided in order to prevent too many startups per hour as mentioned above and to avoid unacceptable variations in the water outlet temperatures during the off periods of the units or during defrosting. According to Fig. 11, the number of on–off cycles decreases for higher tank volumes. With the on–off heat pump system, when the volume of the tank increases, the increased tank thermal losses are compensated by energy saving due to the reduction of cycling losses till a volume between 500 and 600 l. Instead, with the capacity controlled system, heat losses from the tank and cycling losses balance already for a tank size of 300 l. No meaningful reductions can be obtained for both systems for tank volumes exceeding 300 liters that seem the proper size for the heating demand of the considered building.
5.2.
In this work simulations were run considering both on–off and modulating air-to-water heat pump systems in a single family house located in Padua (Italy). The analyses of the annual performances of these systems demonstrated that the higher the
The influence of the storage volume
The effect of the thermal storage size in this comparison was investigated too. Simulations were run keeping the capacity
6.
Conclusions
Fig. 10 – SPF profiles of the on–off heat pump (HP) unit and system for different sizing with respect to the annual peak demand of the on–off heat pump system. Energy savings of the capacity controlled heat pump system with respect to the on–off heat pump system are reported too.
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Fig. 11 – On–off cycles, cycling losses (CL), tank thermal losses (TL) and total energy saving of the modulating unit toward the on–off one as a function of the tank volume.
need for modulation, the higher the advantages of the modulating system over the on–off one. Anyway, these advantages strongly depend on the sizing of both the on–off unit and the thermal storage. The maximum SPF of the on–off system was obtained with the heat pump unit sized to satisfy between 59% and 72% of the building peak load, which allows to meet about 95%–98% of the annual heating demand. One of the main conclusions of this study is that, when comparing a modulating heat pump with an on–off heat pump system, the respective SPFs values are not the only elements to be considered. Other variables are significant, such as the heat pump capacities for a given peak load or the volume of the thermal storage. Finally, cycling losses cannot be neglected when evaluating heat pump annual performance. When an on–off system is sized to face a high fraction of the building peak load, cycling losses might exceed 12% of the whole electric consumption of the system, particularly if the volume of the thermal storage is small. REFERENCES
Adhikari, R.S., Aste, N., Manfren, M., Marini, D., 2012. Energy savings through variable speed compressor heat pump systems. Energy Procedia 14, 1337–1342. Aprea, C., Mastrullo, R., Renno, C., 2006. Experimental analysis of the scroll compressor performances varying its speed. Appl. Therm. Eng. 26, 983–992. Bagarella, G., Lazzarin, R.M., Lamanna, B., 2013. Cycling losses in refrigeration equipment: An experimental evaluation. Int. J. Refrigeration 36 (8), 2111–2118. Bagarella, G., Lazzarin, R., Noro, M., 2016. Annual simulation, energy and economic analysis of hybrid heat pump systems for residential buildings. App. Therm. Eng. doi:10.1016/ j.applthermaleng.2016.01.089. Cecchinato, L., Chiarello, M., Corradi, M., 2010. A simplified method to evaluate the seasonal energy performance of water chillers. Int. J. Therm. Sci. 49, 1776–1786. Cheung, H., Braun, J., 2014. Performance comparison for variablespeed ductless and single-speed ducted residential heat pumps. Int. J. Refrigeration 47 (8), 15–25. Directive 2009/28/EC of the European Parliament, 2009. Directive 2009/28/EC of the European parliament and of the council of 23 April 2009 on the promotion of the use of energy from renewable and amending and subsequently repealing Directives 2001/77/EC and 2003/30/EC.
Emerson-Copeland, 2015, Selection software v.7.11.. (accessed 15.01.2016.). Ente Nazionale di Normazione, 2008, UNI EN 15316-4-3, Method for calculation of system energy requirements and system efficiencies – heat generation systems, thermal solar systems. Ente Nazionale Italiano di Normazione, 2012, UNI EN 14825: air conditioners, liquid chilling packages and heat pumps, with electrically driven compressors, for space heating and cooling – testing and rating at part load conditions and calculation of seasonal performance. Fahlen, P., Karksson, F., 2005, Optimizing and controlling media flows in heat pump systems. Eighth IEA Heat Pump Conference. Las Vegas, USA. Lazzarin, R.M., 2012. Condensing boilers in buildings and plants refurbishment. Energy Build. 47, 61–67. Lee, C.K., 2010. Dynamic performance of ground-source heat pumps fitted with frequency inverters for part-load control. App. Energy 87, 3507–3513. Lemmon, E.W., Huber, M.L., McLinden, M.O., 2013. NIST Standard Reference Database 23: Reference Fluid Thermodynamic and Transport Properties-REFPROP, Version 9.1. Standard Reference Data Program. National Institute of Standards and Technology, Gaithersburg. Madani, H., Ahmadi, N., Claesson, J., Lundqvist, P., 2010, Experimental analysis of a variable capacity heat pump system focusing on the compressor and inverter loss behavior. International Refrigeration and Air Conditioning Conference. Purdue. Madani, H., Claesson, J., Lundqvist, P., 2011a. Capacity control in ground source heat pump systems part I: modeling and simulation. Int. J. Refrigeration 34, 1338–1347. Madani, H., Claesson, J., Lundqvist, P., 2011b. Capacity control in ground source heat pump systems part II: comparative analysis between on/off controlled and variable capacity systems. Int. J. Refrigeration 34, 1934–1942. Madani, H., Claesson, J., Lundqvist, P., 2011c. Variable capacity heat pump systems: analysis at three different levels. Proceeding of 10th IEA Heat Pump Conference, Tokyo, Japan. NIST, National Institute of Standard and Technology, 2003, Simulation models for finned-tube heat exchangers, EVAPCOND.
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Uhlmann, M., Bertsch, S., 2012. Theoretical and experimental investigations of startup and shutdown behavior of residential heat pumps. Int. J. Refrigeration 35, 2138–2149. Zhao, L., Zhao, L.L., Zhang, Q., Ding, G.L., 2003. Theoretical and basic experimental analysis on load adjustment of geothermal heat pump systems. Energy Convers. Manag. 44, 1–9.
Available online at www.sciencedirect.com
ScienceDirect 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 / i j r e f r i g
Sizing strategy of on–off and modulating heat pump systems based on annual energy analysis G. Bagarella *, R. Lazzarin, M. Noro Department of Management and Engineering, University of Padua, Str. lla San Nicola 3, Vicenza, 36100, Italy
A R T I C L E
I N F O
A B S T R A C T
Article history:
In recent years the use of high efficient variable speed heat pumps has spread widely. Though
Received 15 October 2015
several studies were carried out in the past comparing on–off and modulating strategies
Received in revised form 16
in heat pump units, only few works considered the whole heat pump system. Moreover,
February 2016
because of lack of knowledge about cycling losses, most of comparisons between modu-
Accepted 16 February 2016
lating and on–off strategies did not consider the effects of the reduced coefficient of
Available online 23 February 2016
performance of a system at start-up. In this work simulations were carried out considering both on–off and modulating air-to-water heat pump systems in a single family house
Keywords:
located in Italy. The aim of this study is to evaluate how the heat pump sizing, the thermal
Heat pump systems
storage sizing and the cycling losses phenomenon can influence the annual energy perfor-
On–off
mance of both the systems (on–off and modulating) and to investigate when modulating
Modulation
heat pump plants are actually more efficient than on–off.
Variable speed
© 2016 Elsevier Ltd and IIR. All rights reserved.
Sizing strategy
Stratégie de dimensionnement de systèmes de pompe à chaleur tout ou rien et à modulation basée sur une analyse énergétique annuelle Mots clés : Systèmes de pompe à chaleur ; Tout ou rien ; Modulation ; Vitesse variable ; Stratégie de dimensionnement
1.
Introduction
In recent years the use of high efficient modulating heat pumps (equipped with inverter driven variable speed compressors) has spread widely. From the theoretical point of view several issues
* Corresponding author. Tel.: +39 3478037687; Fax: +39 0444 998888. E-mail address: [email protected] (G. Bagarella). http://dx.doi.org/10.1016/j.ijrefrig.2016.02.015 0140-7007/© 2016 Elsevier Ltd and IIR. All rights reserved.
should be considered when a fixed speed heat pump (here referred to also as on–off heat pump) is compared to a variable speed one (referred to also as capacity controlled or modulating heat pump). Peak load in building heating usually takes place only at the outdoor design temperature (e.g. −5 °C), that is for some tens
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Nomenclature BLDC COP CL Kp PID SPF Tc Td Te Text Ti tm,cycle tm,on
tm,off
two TEV TL
brushless direct current motor coefficient of performance cycling losses [kJ] proportional gain of PID control proportional, integral, derivative control seasonal performance factor condensation temperature [°C] derivative time of PID control [s] evaporation temperature [°C] temperature of the external air [°C] integral time of PID control [s] average water temperature at the condenser outlet during the on–off cycle [°C] average water temperature at the condenser outlet during the on period of an on–off unit [°C] average water temperature at the condenser outlet during the off period of an on–off unit [°C] water temperature at the condenser outlet [°C] thermostatic expansion valve tank thermal losses [kJ]
of hours per year. If the selection of heat pump capacity is based on peak load, for most of the time heat pump capacity shall be controlled as it exceeds heating requirements; a fixed speed heat pump controls capacity by switching on–off. Undersizing an on–off heat pump with respect to the peak load lowers control requirements, however without eliminating them. Of course, undersizing heat pump capacity increases the number of operation hours of the auxiliary system (Lazzarin, 2012). As mentioned above, capacity control of a fixed speed heat pump gives rise to a sequence of on–off cycles that lower the efficiency of the unit. A variable speed unit, on the other hand, could run continually, so reducing the on–off frequency and the cycling losses phenomenon (Bagarella et al., 2013).
In principle, a modulating heat pump is more efficient than an on–off unit of the same nominal heating capacity. In fact, the same heating energy during part load operation is given by the modulating heat pump and by the on–off one, but the latter works only for a fraction of time, whereas the former operates continuously. Hence, the average temperature of the secondary fluid, during the on period, must be higher for the on–off unit to allow the due heat exchange. Fig. 1 depicts the typical water temperature profile at the outlet of an on– off heat pump unit. When the heat pump is activated (on period), the water temperature increases. As soon as the water temperature reaches the upper limit (whose value depends on the control settings) the unit is switched-off by the controller. Then, the water temperature decreases, due to the building thermal load. Finally, once the temperature reaches the lower limit (which also depends on the control settings) the heat pump unit is switched on again. In Fig. 1, the outlet water temperature of a modulating unit can be evaluated as the average water temperature of the on–off one (tm,cycle) during the whole period, which is clearly lower than the average water temperature of the on–off unit during the on period (tm,on). This will affect the condensation temperature the same way, leading the modulating heat pump to operate with a lower condensation temperature, then with higher COP (coefficient of performance, ratio of heating provided to work required). From the previous rationale a meaningful increase in thermodynamic efficiency is waited when considering a variable capacity heat pump towards a fixed one. In fact, Aprea et al. (2006) compared an on–off air-water heat pump with a modulating unit and found a 20% electric energy reduction for the latter. Energy savings were mainly attributed to the reduction of the compression ratio during the modulation phase (Aprea et al., 2006). However, a thorough energy analysis of the system shall consider the electricity needs of the inverter that equips a variable speed heat pump. Though several studies (e.g. Aprea et al., 2006, Shao et al., 2004, Zhao et al., 2003) have been carried out in the past comparing on–off and modulating strategies in single heat pump units, only few studies (Adhikari et al., 2012; Cheung and Braun, 2014; Lee, 2010; Madani et al., 2011b) analyzed the whole heat pump system, which includes not only the single heat pump
Fig. 1 – Outlet water temperature profiles in on–off heat pump unit (Fahlen and Karksson, 2005).
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unit, but also the heat source, the heat sink, the heating distribution system, the thermal storage, the liquid pump and the auxiliary heaters, as explained in section 2. Lee (2010), Adhikari et al. (2012) and Cheung and Braun (2014) simulated annual performance of both on–off and capacity controlled heat pump systems in different climates and concluded that a capacity controlled system can achieve considerable energy savings compared to an on–off one, especially in mild climates. Anyway, these studies did not consider, for a single climate, the possibility of sizing the on–off system to satisfy different fractions of the building peak load. Does the energy saving of a modulating heat pump system (for a given climate) depend on the size of the benchmark on–off system? Only Madani et al. (2011b) raised this issue, simulating the annual operation of both on–off and modulating geothermal heat pumps with the typical thermal load of a single house located in Stockholm (Sweden). Assuming as auxiliary system an electric heater, they concluded that the modulating heat pump led to a better performance only when the on–off heat pump was sized to cover less than 65% of the heating peak demand. On the other hand, when the on–off heat pump presented a higher capacity, no energy savings were found because benefits deriving from a lower mean compression ratio allowed by the modulating heat pump were completely compensated by the energy requests of both inverter and pumps. In this paper, as in Madani et al. (2011b), several simulations were carried out considering on–off and modulating heat pump systems (with different capacities) with the typical thermal loads of a single family house located in Padua (NorthEast of Italy). The main difference between this study and the above cited (Madani et al., 2011b) is in the heat pump source, since outside air (instead of ground) is here considered. Heat pump capacity is strongly influenced by the heat source temperature and the annual variations of the air temperature is much higher than the maximum variation of the ground temperature (e.g. 3 K was the maximum annual ground temperature variation considered by Madani et al., 2011a). Moreover, when the air temperature decreases with a reduction of heat pump capacity, the building thermal load usually increases. There-
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fore, the heat-pump capacity modulation ability assumes a crucial importance in air-to-water applications, whereas it represented a minor problem in the ground source system studied by Madani et al. (2011a, 2011b). Another major difference is in the choice of the auxiliary heater for the heat pump system. Whereas in Sweden electric heaters are common, probably due to favorable tariffs, gas boilers are used in Italy. Therefore, a condensing boiler (fueled by natural gas) was considered. Finally, probably because of lack of knowledge about the cycling losses phenomenon, most of the comparisons between modulating and on–off strategies found in literature (including the study proposed by Madani et al., 2011b) do not consider the effects of the reduced COP of a system at start-up. In this paper cycling losses penalizations have been duly considered basing on a previous work specifically devoted to this computation (Bagarella et al., 2013).
2.
Methodology
To simulate the annual behavior of both on–off and variable capacity heat pumps the system was firstly modeled. Fig. 2 shows a schematic overview of the modeled heat pump system. The air-to-water heat pump unit supplies heat to the building by means of a hot water storage tank and is modeled by linking together the sub-models of condenser, expansion valve, evaporator and compressor as in Madani et al. (2011b). Some technical data of the considered heat pump unit are summarized in Table 1. REFPROP 9.0 (Lemmon et al., 2013) was used to estimate the refrigerant properties in each operating condition. Heat transfer coefficients of both evaporator and condenser were set according to results obtained by tests carried out with an equivalent experimental unit as in Madani et al. (2011c). Superheating and subcooling were considered constant as in Cecchinato et al. (2010). The global electric consumption of the system was calculated summing up compressor, inverter, pump and fan require-
Fig. 2 – Schematic overview of the heat pump system.
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Table 1 – Heat pump unit data. Component Condenser
Evaporator
Refrigerant Nominal heating capacity (Text = −5 °C) Expansion valve Compressor
Table 2 – Building data. Information
Plate heat exchanger Heat exchange surface: 0.79 m2 Plates: 20 Finned coil Wavy fins Tube length: 920 mm Number of columns: 48 Number of rows: 2 Fin thickness: 0.15 mm Fin pitch: 1.95 mm Tube spacing: 25.4 mm R410A 5.9 kW TEV Kv = 0.066 m3 h−1 BLDC Twin-rotary Displacement: 1.3 dm3
ments. The compressor consumption was estimated basing on manufacturers’ data for each operating conditions. Pump and fan consumption (0.03 kW and 0.06 kW respectively) was assumed constant. The inverter efficiency was considered in a range between 88% and 95%, depending on the compressor rotation speed, as specified by the manufacturer. Cycling losses were measured by specific experimental tests as described in Bagarella et al. (2013) and results were included in the model. In that study, details about how the number of hourly start-ups affects the percentage efficiency loss of the unit are reported and deeply discussed. One of the main conclusions of that study is that the cycling losses phenomenon strongly depends on the expansion valve used. If a thermostatic expansion valve is used, the COP of the unit at the start-up is strongly reduced. This behavior is not due to a higher electric power absorbed by the compressor, but rather to the poor heating effect during the starting transient. Experimental tests demonstrated that, once the unit is switched on, steady state conditions are attained only after 100–120 seconds. During this transient period, the average COP is approximately 50% lower than the steady state value. Basing on this consideration, the algorithm proposed by Madani et al. (2011b) was modified, and the unit was forced to operate with a reduced COP (50% lower than the one calculated through Madani’s model) during each start-up period (110 seconds). The compressor manufacturer claims a maximum number of hourly on–off cycles up to 12. Nonetheless, the maximum number of startups here considered is 5, as more cycles would excessively reduce the compressor life. A simplified method to account for frost formation and defrost cycle consumption was also implemented in the model. Details about this simplified model (i.e. information regarding main assumptions, software used to roughly account for frost formation and model validation) are reported in Bagarella et al. (2016). The average fin temperature was estimated using software EVAP-COND developed by NIST (NIST, National Institute of Standard and Technology, 2003). When fin temperature gets conditions that favor frost formation (fin
Parameter Building type Building volume Building height Vertical structures transmittance Horizontal structures transmittance Floor transmittance Windows transmittance Ventilation rate Set point indoor temperature Peak load (Text = −5 °C) Annual thermal energy demand
Information Single family house (2 floors) 363.5 m3 5.5 m 0.28 W m−2 K−1 0.24 W m−2 K−1 0.27 W m−2 K−1 1.60 W m−2 K−1 109 m3 h−1 20 °C 6.1 kW 5757 kWh
temperature below 0 °C and below outdoor air dew point), frost starts building up. Once the frost has reached its upper bound (average frost thickness 0.7 mm) defrost (by electric heaters) is activated. The model of the heat pump unit was validated comparing its behavior against those of the real modulating one. In particular, a four hour test in a climatic chamber was carried out (main characteristics of the real unit are described in Table 1) according to test conditions suggested by Ente Nazionale Italiano di Normazione (2012) (National Standard Agency) for an average European climate. During the test, for each partload condition suggested by Ente Nazionale Italiano di Normazione (2012), the following quantities were measured and compared with those calculated by the model (with the unit in steady-state condition): evaporation and condensation temperatures, compressor frequency and thermal power. At the end of the test, the electric energy absorbed by the capacity controlled unit (5.9 kWh) was compared to the electric energy calculated by the dynamic model (5.7 kWh). The same procedure was followed to validate the on–off unit (the compressor of the real capacity controlled unit was forced to operate at its maximum speed during the on periods, regulating its capacity by switching the compressor on–off to produce the same thermal energy as the capacity controlled unit). The electric energy absorbed by the real on–off unit (7.5 kWh) was 5% higher than the electric energy calculated by the dynamic model (7.1 kWh). The building (Table 2) was modeled in TRNSYS 16.0, estimating the thermal loads under annual climatic data (Test Reference Year, TRY, 1986) of Padua (Italy). The design temperature is −5 °C and heating degree days are 2381. The thermal properties of the building envelope are those compliant to the present National Regulations for new buildings’ standards. The thermal load of the building and the thermal capacity of the heat pump unit were used to continuously calculate the temperature of the hot water tank (Fig. 2). The fully-mixed thermal storage decouples the heat produced by the heat pump from the heat required by the building. The storage protects the heat pump unit from an excessive number of on–off cycles and ensures the thermal comfort inside the building avoiding excessive variations in the air temperature during defrosts. A low temperature radiant heating distribution system was considered. The set-point of the hot tank was set at 35 °C.
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Fig. 3 – Compressor isentropic efficiency at different frequencies and evaporation (Te) and condensation (Tc) temperatures.
The control unit allows the heat pump to work either with on–off cycles or with a capacity control algorithm. The control variable was the temperature of the thermal storage. In the case of on–off strategy a 2 °C dead band around the set-point of the water tank was used, while in the capacity controlled unit a PID algorithm was used to set the compressor speed. When the heating capacity of the heat pump unit fails to cover the heating demand, a condensing boiler auxiliary heater is activated. The condensing boiler (fueled by natural gas) is modeled through its thermal efficiency, which is assumed a function of the water temperature at the boiler inlet. The thermal efficiency profile suggested by Lazzarin (2012) was used. As in Table 1, the nominal thermal capacity of the real heat pump is 5.9 kW (when Text = −5 °C). According to this validated model, other heat pumps of different nominal capacity were simulated as explained in section 5.1. Climatic conditions have two effects on the system: lower air temperatures mean lower evaporation temperatures (consequently lower COPs) and, at the same time, higher building heating demand. The global algorithm of the heat pump system was implemented in LABVIEW 8.5 environment by linking together the sub-models of the elements following the approach in Fig. 2.
3. COP variations and modulation capacity of the variable speed unit A first application of the model of both heat pump units (on– off band modulating) and heat pump systems was devoted to study the capacity controlled unit performance. The influence of outside air temperature and heating demand on the heat pump COP was investigated. For a better understanding of the results, compressor isentropic efficiencies (calculated basing on manufacturer data) at different frequencies are reported in Fig. 3. The maximum isentropic efficiency of the compressor is obtained for frequency from 50 Hz to 70 Hz. Similar results were found by Madani et al. (2010). As waited, the isentropic efficiency lowers for higher pressures ratios. However, for very low condensation temperatures (e.g. Tc = 28 °C), that is for a low pressure ratio, the isentropic efficiency lowers as well, probably because the used twinrotary compressor was not optimized for really low pressures ratios. Fig. 4 represents the COP of the modulating unit as a function of the thermal load of the building for five different external air temperatures. The capacity at the maximum frequency
Fig. 4 – Heat pump unit COPs at different temperatures of the external air (Text, which is equal to the evaporator air inlet temperature) and for different thermal loads.
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Fig. 5 – Behavior of both high reactive and high stable compressor PID controls.
(120 Hz) lowers for decreasing outside air temperature. The maximum COP point shifts toward lower capacities too because the compressor is optimized for a given frequency (e.g. 50 Hz), and the lower Text, the lower the thermal capacity at which this frequency is reached. The most important feature of the curves in Fig. 4 is the significant decrease of the COP (after the maximum value) moving toward higher capacities. The modulation ability of an inverter driven heat pump unit should not be overrated. When the nominal frequency of the compressor is 120 Hz, its velocity can be reduced to 20– 30 Hz. In fact, when the compressor frequency is lower than this minimum value, a strong reduction in volumetric efficiency takes place. As a matter of fact, at low rotation speeds the lubrication does not assure a good tightness against the return of refrigerant to the suction. This problem affects all the most widespread small size compressor models. Scroll compressors allow a sensible advantage compared to rotary and twin-rotary compressors, since lower minimum velocities (20 Hz, instead of 30 Hz) are sometimes claimed by manufacturers (Emerson-Copeland, 2015; Siam Compressor Industry, 2010). Moreover, the electric motor efficiency usually reduces for lower frequencies. Furthermore, the higher electric consumption at low rotation speed is converted into thermal power and then transferred to the refrigerant. Hence, the modulation ability is far from being proportional to the reduction in rotational speed of the compressor. This ability is further decreased when the outside temperature is higher, since the heat pump capacity increases just when an effective modulation would be required by the decrease of building thermal load. In addition to the minimum compressor frequency, also the PID parameters of the compressor controller affect the modulating heat pump minimum capacity. As depicted in Fig. 5, the higher the reactivity of the controller (how fast the controller reacts to a variation of the controlled variable, which is the water temperature at the condenser inlet), the higher the overshoot of the compressor frequency (difference between the final and the minimum value of the process variable, which is the compressor frequency), thus the higher the minimum thermal load which can be satisfied by the heat pump unit avoiding the compressor frequency to fall below 30 Hz. In fact, when the heating load decreases, if the reactivity of the controller is too high the
compressor frequency rapidly decreases too and it may fall below the minimum compressor frequency (specified by the manufacturer). In this case, the heat pump is switched off, as a protection of the compressor (especially if this operating condition lasts for several minutes). This usually happens when the proportional gain (Kp) is set to a high value and the integral time (Ti) is set to a low value. On the other hand, the higher the stability of the control, the longer the necessary time to ensure that the heat pump capacity matches the load. Table 3 reports the capacity ratio (the ratio of minimum capacity to the maximum capacity) of the considered compressor (minimum frequency set to 30 Hz) when the parameters of the PID controller are set in order to let the heat pump operate in reasonably stable conditions. The reader should be aware that, while the capacity ratios reported in Table 3 were obtained through a dynamic simulation of the heat pump, the profiles in Fig. 5 were proposed with the only purpose of representing high reactive and high stable controls. To exemplify how the PID parameter setting affects the reactivity of the regulation, coherently with the real settings used in the simulations (Kp = 10, Ti = 180 s; Td = 0 s), the high reactive and the high stable controls in Fig. 5 might be respectively obtained with the following parameters: Kp = 20, Ti = 40 s, Td = 0 s and Kp = 5,Ti = 500 s, Td = 0 s.
4.
Comparison during one whole year
Dynamic simulations were carried out with a 15 minute time step to compare the performance of the plant equipped with
Table 3 – Capacity ratio of the heat pump unit at different temperatures of the external air. Text (°C) 16.0 12.0 7.0 2.0 −7.0
Maximum capacity (kW)
Minimum capacity (kW)
Capacity ratio%
8.2 8.0 7.2 6.7 5.8
4.4 4.0 3.3 2.9 2.2
53.8% 50.0% 46.5% 42.9% 38.2%
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Fig. 6 – Bivalent temperature of the considered heat pump (HP).
on–off or modulating air-to-water heat pump. Firstly a 150 liter thermal storage for both on–off and modulating heat pump was considered. In the first simulation the two heat pump units were sized to cover the 91.4% of the maximum thermal load required for heating up the house. This size of both heat pumps corresponds to a −4 °C bivalent temperature, as reported in Fig. 6. From a theoretical perspective the bivalent temperature may be defined as the minimum temperature at which the thermal capacity of the heat pump matches completely the building thermal load (Ente Nazionale Italiano di Normazione, 2012). When the outside air temperature is lower than the bivalent temperature an auxiliary heater is thus required. On the other hand, when the outside air temperature is higher than (or equal to) the bivalent temperature, the heat pump operates either modulating or in on–off, as the nominal heat pump capacity overrates the load. Fig. 7 shows the different annual profiles of the monthly seasonal performance factor (SPF) of both capacity controlled and on–off heat pump systems. The seasonal performance factor of a heat pump system is defined as the ratio between the seasonal thermal energy delivered by the plant (including auxiliary heaters) and the energy input required by the whole heat pump system (Directive2009/28/EC of the European
Parliament, 2009). For most of the time the heating load is less than the heating capacity of the heat pumps. Under these conditions the capacity controlled heat pump unit can work with higher SPF than the on–off one. The differences in the SPF values of the two systems are higher in spring and autumn, when is higher the need for modulation, as the building load is lower. The capacity controlled unit can work continuously reducing the compressor frequency, thus operating with lower pressure ratios than the on–off unit, consequently with less compressor specific work. However, the differences in the SPF values are lower when the heating loads approach the nominal capacity of the heat pump units. Then the capacity controlled heat pump unit operates at full capacity, without the benefits due to modulation. Fig. 7 also reports the cumulative number of on–off cycles for both heat pump systems. Though the capacity controlled system leads to a drastic reduction of on–off cycles, its modulation capacity decreases when the outdoor air temperature is higher. This phenomenon is confirmed by the profile of the cumulative on–off cycles curves, whose slope increases in spring and autumn with respect to the coldest periods. Fig. 8 illustrates a breakdown of the energy balance of the system. The electric consumption of the variable speed unit
Fig. 7 – Annual profiles of the monthly SPF and cumulative on–off cycles of both capacity controlled and on–off heat pump systems. Monthly thermal energy needs and monthly maximum loads are reported too.
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Fig. 8 – Absolute and percentage breakdown of energy required by the system with both on–off and capacity controlled heat pump systems.
is 1662 kWh, 13.2% less than the electric energy absorbed by the on–off heat pump. The pump and the fan consumption are higher in the capacity controlled system since they operate for a longer time. Moreover, the inverter energy cost must be accounted for in the capacity controlled unit. In the specific conditions of the simulation, the energy cost of defrosting is slightly higher for the capacity controlled unit. This behavior could be attributed to the fact that the modulating unit operates for a longer time during the heating season. Last but not least, the capacity controlled heat pump system strongly limits the total number of on–off cycles, as described in Fig. 7, and so it definitely reduces cycling losses. Fig. 8 suggests that, if annual performances of the two systems are compared, cycling losses cannot be neglected. In the present simulation the effect has been estimated to be 3.4% and 8.3% of the total electric consumption respectively for the modulating and the on–off heat pump systems. These values are higher than the 2.1% efficiency loss reported by Uhlmann and Bertsch (2012), who considered a maximum number of 3 hourly start-ups. The main difference in the two simulations can be identified in the choice of a water tank whose volume is finalized to limit to 5 the maximum number of hourly on–off cycles in this work. In fact, real heat pump systems are usually designed as to avoid the compressor to operate with an on– off frequency too high. The reason is that, even if the effects of the cycling losses phenomenon are often ignored, an on– off frequency too high shortens compressor life. In real
applications, even if the experienced designer has sized the volume of the water storage to avoid a maximum number of hourly on–off cycles higher than 5, a sixth hourly activation may sometimes happen. However, this occurrence is rather rare so that the effects on the compressor life are negligible.
5. Analysis of main variables influencing the seasonal performance 5.1.
The influence of the unit size
To analyze how the sizing of the on–off heat pump can influence system performance, several simulations were run varying the nominal heating capacity of the heat pump, meeting a fraction from 74.3% to 99.8% of the annual heating load. On the contrary, the nominal capacity of the inverter driven heat pump was at first kept constant (5.9 kW). As in Fig. 9, the higher the heating capacity the lower the bivalent temperature (thus the lower the auxiliary energy required). Table 4 reports the annual number of on–off cycles of the on–off heat pump for different fractions of the satisfied demand. The number strongly increases with that fraction and so the cycling losses. Accordingly, an SPF reduction of the unit is observed. On the other hand, the lower the heating capacity of the heat pump considered, the higher the auxiliary energy used.
Fig. 9 – Bivalent temperature depending on the heating capacity of the heat pump unit (HP).
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Table 4 – Total energy, on–off cycles, SPF of the unit, SPF of the system and percentage energy savings of the capacity controlled system for different covering percentages of annual heating demand (or peak load) of the on–off heat pump system. Annual heating demand covering % Peak load covering % On–off cycles Auxiliary energy (kWh) Cycling losses (kWh) Total energy consumption (kWh) SPF on–off HP unit SPF on–off HP system Inverter HP savings
74.3% 33.5% 1237 666 43 1994 3.34 2.89 16.2%
81.7% 38.6% 1725 473 60 1956 3.27 2.95 14.6%
87.% 43.8% 2233 321 78 1931 3.22 2.99 13.4%
91.9% 54.1% 2700 210 95 1914 3.18 3.01 12.7%
94.9% 59.2% 3140 132 110 1909 3.14 3.02 12.4%
97.1% 65.7% 3542 75 124 1906 3.12 3.02 12.3%
98.4% 72.1% 3896 41 136 1909 3.09 3.02 12.5%
99.7% 91.4% 4574 9 160 1924 3.05 3.00 13.1%
99.8% 94.0% 4837 5 169 1932 3.04 2.98 13.5%
Thus the weight of SPF of the auxiliary heater over the SPF of the whole system is growing. As the auxiliary heater performance is always lower than the heat pump, in Table 4 and Fig. 10, the SPF of the on–off heat pump system is always lower than the SPF of the on–off heat pump unit. The highest value of the system SPF was obtained in this study for a fraction of the annual heating demand satisfied from 95% to 98%, which corresponds (in the specific climatic conditions of Padua) respectively to a 59%–72% nominal capacity with respect to the peak thermal load. Fig. 10 and the last row of Table 4 indicate that the energy savings of the capacity controlled heat pump, with respect to the on–off system, always exceeds 12%. Whereas an oversizing of the on–off heat pump, with regard to the above mentioned optimum capacity, slightly penalizes the system, an undersizing is more severe for the low efficiency contribution of the auxiliary. A further study was devoted to determine the best sizing of the inverter driven heat pump. However, every further reduction of its capacity with respect to the above simulation revealed a worse performance. The main reason is that a reduction of cycling losses for lower capacities does not compensate the penalty of the lower performance of the increasing auxiliary operation.
of the heat pump units constant and varying the volume of the tank for both the considered systems. In the simulations the thermal losses of the storage (a 2 °C dead band around the set point was used) have been considered according to Ente Nazionale di Normazione (2008). A minimum 100 liter water storage volume was provided in order to prevent too many startups per hour as mentioned above and to avoid unacceptable variations in the water outlet temperatures during the off periods of the units or during defrosting. According to Fig. 11, the number of on–off cycles decreases for higher tank volumes. With the on–off heat pump system, when the volume of the tank increases, the increased tank thermal losses are compensated by energy saving due to the reduction of cycling losses till a volume between 500 and 600 l. Instead, with the capacity controlled system, heat losses from the tank and cycling losses balance already for a tank size of 300 l. No meaningful reductions can be obtained for both systems for tank volumes exceeding 300 liters that seem the proper size for the heating demand of the considered building.
5.2.
In this work simulations were run considering both on–off and modulating air-to-water heat pump systems in a single family house located in Padua (Italy). The analyses of the annual performances of these systems demonstrated that the higher the
The influence of the storage volume
The effect of the thermal storage size in this comparison was investigated too. Simulations were run keeping the capacity
6.
Conclusions
Fig. 10 – SPF profiles of the on–off heat pump (HP) unit and system for different sizing with respect to the annual peak demand of the on–off heat pump system. Energy savings of the capacity controlled heat pump system with respect to the on–off heat pump system are reported too.
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Fig. 11 – On–off cycles, cycling losses (CL), tank thermal losses (TL) and total energy saving of the modulating unit toward the on–off one as a function of the tank volume.
need for modulation, the higher the advantages of the modulating system over the on–off one. Anyway, these advantages strongly depend on the sizing of both the on–off unit and the thermal storage. The maximum SPF of the on–off system was obtained with the heat pump unit sized to satisfy between 59% and 72% of the building peak load, which allows to meet about 95%–98% of the annual heating demand. One of the main conclusions of this study is that, when comparing a modulating heat pump with an on–off heat pump system, the respective SPFs values are not the only elements to be considered. Other variables are significant, such as the heat pump capacities for a given peak load or the volume of the thermal storage. Finally, cycling losses cannot be neglected when evaluating heat pump annual performance. When an on–off system is sized to face a high fraction of the building peak load, cycling losses might exceed 12% of the whole electric consumption of the system, particularly if the volume of the thermal storage is small. REFERENCES
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