The results presented in this study are based on measured data of the cooling period from June till September in 2016 and the following heating period from January till March in 2017. During the considered periods the test facility was operated from 7 am to 10 pm every day of the week. Transition periods in spring and fall are not considered in this study, because these periods are not suitable to analyze strengths and weaknesses of the system as a reason of the climate conditions in northern Germany. Volume flow of supply air was controlled to be constant in the range of ; mass flow rates of supply and extract air were controlled to be equal. Set point of sup water content is for dehumidification mode. The following evaluation is subdivided into four parts. First, the system is evaluated regarding relevant performance parameters of summer and winter operation and the performance of the geothermal system is evaluated. Afterwards, thermal comfort within the air conditioned space is analyzed. Finally, the investigated system is compared to different reference systems focusing electrical and thermal energy demands. System performance is evaluated separately for summer and winter operation in general for this study in order to show strengths and weaknesses for each operation mode.
3.1. Performance Evaluation of the Air Conditioning System
The following evaluation of system performance is based on measured data during the investigated periods. First, electrical power demand is considered. Electrical power demand of the entire system is in the range of 770–900 W
el during summer operation. The fans account for about 81–95% of this power demand, whereas the remaining part is divided equally between other auxiliary energies (e.g., drive of the wheels, circulation pumps). The electrical power demand of the GCHP has to be considered additionally during winter operation (
). Indexing within the following equations is according to
Figure 2.
To evaluate the air handling unit for the considered periods, electrical and thermal COP values are used. These performance indicators are defined according to [
33]:
All figures shown in the following rely on steady-state operation conditions. Measured data were selected for steady state operation 15 minutes after the last changes made by system control. The electrical and thermal COP of the air handling unit during summer operation at dehumidification mode are shown in
Figure 4 in dependence of regeneration air temperature. A strong dependence between
and regeneration air temperature is visible from the plot in
Figure 4a. Due to the fact that the electrical energy demand of the AHU is nearly independent of the outside air conditions,
is increased with increasing regeneration air temperature. Increase in regeneration air temperature is a result of increasing water content and/or temperature of outside air within dehumidification mode. With respect to
Figure 4b, desired supply air temperature cannot be maintained at high outside air temperature and water content, causing an increase in supply air enthalpy, respectively. This causes flattening of the curve for
with increasing regeneration air temperature. Due to its definition, negative values of
occur at low outside air temperatures when dehumidification of supply air is necessary whereas cooling is not required.
At values of regeneration air temperature below 35 °C, the increase of AHU thermal COP with increasing regeneration air temperature is much steeper compared to the slope of at higher regeneration air temperature. This characteristic results from the mathematical definition of as presented in Equation (2). Negative values of occur at low regeneration air temperature when dehumidification is still necessary and reheating of supply air is required at the same time. For higher regeneration air temperatures above 35 °C, thermal COP keeps nearly constant at with decreasing fluctuations for increasing regeneration air temperature.
For the investigated heating period, performance indicators in form of
and
are shown in
Figure 5. To evaluate the performance of the air handling unit during winter mode, the characteristics of
and
depending on outside air temperature are presented for EW mode and HRW mode.
As shown in
Figure 5a, the electrical COP of the air handling unit is decreasing with nearly constant gradient for increasing outside air temperature in both operation modes. With respect to Equation (1), this is a result of decreasing nominator with increasing outside air temperature, whereas the denominator keeps constant in good approximation. Generally, water content of outside air is sufficient regarding indoor air comfort limits at higher outside air temperature level. Thus, HRW mode is just occurring at outside air temperatures above 4 °C. Due to similar electrical energy demand of the wheels as well as similar pressure drop across these components, the resulting gradients of
are similar. The dependence of thermal COP on outside air temperature for the air handling unit during winter operation is shown in
Figure 5b. In spite of increasing fluctuations with increasing outside air temperature,
is independent of outside air temperature and not dropping below
. This is a result of its definition, as presented in Equation (2), with similar characteristic of nominator and denominator for the underlying boundary conditions.
To further evaluate desiccant assisted dehumidification and enthalpy recovery in detail, mode-specific key figures have to be defined. In order to take the fact into account that the desiccant wheel is used for active dehumidification with a regeneration air heater, dehumidification efficiency is defined as follows:
with
according to Equation (3) and is therefore equivalent to the definition of latent COP. An average dehumidification efficiency of
was achieved for the considered cooling period. This result indicates that more latent thermal power was absorbed within the desiccant wheel compared to the required thermal power to run the regeneration air heater. In general, regenerative heat exchange within the air handling unit improves dehumidification efficiency by preheating extract air for regenerating the desiccant wheel. Due to the fact that the wheel is used for passive enthalpy recovery during winter, moisture recovery efficiency is expressed by:
An average moisture recovery efficiency of was achieved for the enthalpy wheel during the investigated winter period. Thus, an increase of was achieved for sup humidity ratio on average. Maximum values of moisture recovery were close to for the underlying boundary conditions.
3.2. Performance Evaluation of the Geothermal System
Performance evaluation of the overall geothermal system is structured into three parts. First, soil temperature and soil energy balance are considered. Afterwards, energy transfer at the BHE is investigated and finally, GCHP performance is evaluated.
Temperature profiles of BHE and the considered reference BHE are shown in
Figure 6a for the period of around one year including summer and winter operation. The soil temperature 15 m below ground surface was found to be independent of seasonal related temperature fluctuations during previous investigations. Thus, the plots result from temperature averaging below 15 m for both, BHE and reference BHE. Depending on the season and operation mode of the air conditioning system, the the soil around the BHE is significantly influenced with dynamic temperature profile during cooling and heating mode. Cooling peak loads occurring in summer operation lead to maximum soil temperatures above 18 °C. This temperature level is crucial with respect to keep the desired indoor air temperature level below 25.5 °C. Regardless of such peak loads that only occurred at a few days of the considered cooling period, the soil temperature was kept within a sufficient temperature range in terms of cooling purposes. During winter operation, the soil temperature is less fluctuating compared to its use as heat sink. The lowest soil temperature was 4.5 °C that occurred during the coldest period in the beginning of February. Using the soil for heating, the soil temperature decrease is crucial to operate the GCHP system efficiently. This dependence is further analyzed later on in this subsection. Average temperature level of the undisturbed soil at the reference BHE was at 9.8 °C. Occurring temperature fluctuations were within the corresponding uncertainty of temperature measurement.
Balancing input and output of thermal energy during summer and winter operation and natural regeneration of the soil, an equalized energy balance of the soil was achieved. Input and output of thermal energy at the BHE are balanced with a remaining annual difference of
, 9% respectively, as shown in
Figure 6b. This difference is within measurement uncertainty of the corresponding energy values. With respect to these results an efficient long-term operation of the geothermal system can be predicted. Nevertheless, an ongoing long-term monitoring of the geothermal system is essential, especially when large scale geothermal systems with several BHE influencing each other are considered.
In terms of further investigating energetic performance of the geothermal system, month and period specific performance indicators are presented in
Figure 7. For both periods the amount of energy (
) and heat flow (
) transferred at the BHE as well as resulting performance values are shown. Key figures as defined in [
33] are used to evaluate BHE performance. With respect to the considered periods, Monthly Performance Factor (MPF) and Seasonal Performance Factor (SPF) are used as presented in Equation (6):
The denominator includes the electrical energy demand of the BHE circulation pump. Decreasing MPF values occurred over each period as a result of changing temperature level of the soil surrounding the BHE by charging or discharging energy in form of heat. The amount of thermal energy, heat flow and resulting MPF show the same relationship for both periods with one exception. Even though the month of July shows the largest amount of thermal energy transferred to the soil, the highest MPF was achieved in June with
, see
Figure 7a. This was primarily caused by lower temperature increase of the soil during June. Evaluating the entire cooling period, a seasonal performance of
was achieved, indicating a high efficiency of the geothermal heat sink. The same holds true for the winter period with a resulting seasonal performance of
, even though the value is around 28% lower compared to summer mode. The reason for this difference is related to the required volume flow of heat transfer medium to supply the evaporator of the heat pump that is generally higher than volume flows of heat transfer medium for natural cooling in summer.
To further analyze GCHP performance, its electrical COP, see Equation (7), is investigated in more detail in terms of available and required temperature levels as shown in
Figure 8:
BHE outlet temperatures were mostly pooled in the range of
during steady state GCHP operation with a range of
, as shown in
Figure 8a. A trend of increasing
with increasing BHE outlet temperature is visible. This is an effect of lower required temperature lift within GCHP process that comes along with reduced power level required to run the compressor. Taking the required temperature lift as a further indicator of GCHP load into account, a slight dependence on GCHP performance can be deduced from
Figure 8b. The temperature lift
is defined as temperature difference between condenser outlet and evaporator inlet. Values above
that were required to supply UHS caused GCHP performance lower than 3 with decreasing trend curve. Taking the overall heating period into account, SPF of the GCHP system can be determined equivalent to Equation (6) by integrating thermal and electrical powers from Equation (7):
For the considered heating period was achieved. Compared to GCHP systems state of the art with , performance of the investigated system relying on a reciprocating compressor shows potential for improvement. Nevertheless, this system is robust against fluctuating temperature levels.
3.3. Evaluation of Thermal Comfort
The quality of indoor air conditions is a result of the performance and operation strategy of the overall system. To provide an overview of comfort conditions for the investigated periods, outside and room air conditions during system operation are shown in
Figure 9a and
Figure 10a with simplified comfort areas according to DIN EN 15251 [
38]. Comfort areas of category I and II are defined to ensure less than 6% (cat. I) and 10% (cat. II) of occupants being dissatisfied with the present indoor air conditions. In order to exclude start-up effects, the first hour of system operation is not considered.
For the investigated cooling period, indoor air conditions according to cat. I were maintained for 55% and cat. II was maintained for over 96% of operation time, respectively. The reasons for the remaining violations were different for cat. I and cat. II. Cat. I was violated primarily due to too high indoor air humidity, whereas too high indoor air temperatures caused violations of cat. II.
In order to further analyze the reasons for overheating,
Figure 9b shows the dependence of room air temperature
on water outlet temperature
of the BHE. Outlet temperature of the BHE is similar to the cooling ceilings’ inlet temperature. With good approximation, a linear increase of indoor air temperature with increasing BHE outlet temperature above 18 °C is visible. Maximum room air temperatures were right above 26 °C at maximum BHE outlet temperature of
. This is an effect of peak loads that could not be covered by the geothermal heat sink due to its limited capacity and little controllable thermal power output.
According to the German Meteorological Service, the investigated heating period can be classified as moderate. The average oda temperature in January was 1.67 °C at an average water content of
. The following months were characterized by higher oda temperature and water content. As shown in
Figure 10a, nearly 100% of oda conditions were outside comfort area according to cat. I and II. Days with oda conditions within the desired comfort area just occurred during springtime period at the end of March.
Around 67% of indoor air conditions satisfied the requirements according to cat. I. Cat. II was maintained for 75% of system operation time, respectively. The remaining violations were primarily caused by too low indoor air temperature. As a result of the control strategy for the underfloor heating system (UHS) and insufficient internal loads, 85% of indoor air temperatures below the desired temperature level occurred during the first four hours of daily system operation. A similar characteristic applied to indoor air water content. The level of sup and indoor air water content is mainly depending on oda water content if constant internal latent loads are provided. Thus, 75% of indoor air conditions characterized by insufficient level of water content occurred at oda water contents below . Without humidification of supply air, over 60% of room air conditions would have been outside of cat. I. Over-humidification of room air did not occur during the entire heating period.
Figure 10b shows the dependence of indoor air temperature
on UHS inlet temperature
. A trend of increasing fluctuation of indoor air temperature with increasing UHS inlet temperature is noticable, especially for
. This is an effect of insufficient thermal power supply during the first hours after system start-up at low oda temperatures. Taking this initial situation, improvement of the control strategy for the UHS is required in terms of limiting periods at insufficient thermal comfort. The GCHP system was not operated between 10 pm and 7 am, while underfloor heating was provided 24 hours a day. Providing and storing sufficient amounts of thermal energy can be achieved by operating the GCHP system additionally during night. Resulting shorter regeneration periods of the soil have to be considered for the proposed operation strategy focusing on increase in thermal comfort.
Table 3 lists the relative shares of operation modes for both periods for the given outside air conditions and the implemented control strategies. The significant amount of DW or EW mode represents the importance of desiccant assisted air conditioning to provide a high level of indoor air conditions throughout the year.
3.4. System Comparison
The investigated system is compared to different reference systems in order to further evaluate its energetic performance. These reference systems are designed as mathematical simulation models relying on measured data of the investigated system (DW-GEO). Thus, the thermodynamic processes of the considered reference systems were modeled relying on simplified thermodynamic relations and required assumptions. Measurement data of the investigated system were used whenever similar air states are expected for the reference systems in comparison to DW-GEO. In order to ensure a fair system comparison, oda and sup conditions as well as sup mass flow rate and room air conditions were assumed to be equal with one exception that is explained later on. Air dehumidification as well as air humidification within the reference systems are considered according to the equivalent operation modes of the investigated system. In each operation mode, summer and winter, two different reference systems were designed. During summer operation, reference systems relying on an electrical powered vapor compression chiller are considered. Humidifying processes with adiabatic or isothermal air humidification are considered for the reference systems regarding winter operation. The individual characteristics of each reference system are summarized in
Table 4.
Supply air temperature of the conventional reference system (DP-VC) was set to 18 °C in order to ensure a fair system comparison. Changes in thermal load discharge by air and cooling ceilings was considered. The second reference system for the cooling period is a hybrid air conditioning system (DW-VC); the BHE is replaced by a vapor compression chiller, while the rest of the investigated system remains unchanged. The specific layout of corresponding air handling units for the reference systems DP-VC and AH-GEO/IH-GEO is shown in
Figure 11. Highlighted components are specific for the designated reference systems, whereas the rest of the air handling unit remains the same for these systems.
Figure 12 shows the results of the system comparison for the considered periods. Legend entry “supply” includes electricity demands of the compression chiller (DP-VC, DW-VC) or rather the BHE circulation pump (DW-GEO) during summer operation, see
Figure 12a. During winter operation, shown in
Figure 12b, it includes the heat pump and corresponding auxiliary energies, respectively. Electricity demand of the AHU and circulation pumps of the hydraulic circuits are summarized to category “air treatment and distribution” for both periods. Electrical and thermal energy demands are related to the entire periods under consideration. All values of electrical and thermal energy demands shown in
Figure 12 are also listed in
Table 5 for more transparency.
As shown in
Figure 12a, the investigated system boasts significant reductions in electrical energy demand compared to the reference systems during summer operation. The savings sum up to 50% compared to the conventional system (DP-VC) and 34% compared to the hybrid system (DW-VC). These savings were primarily caused by the chiller unit (black column) that required 61% of the total electrical energy demand for the system DP-VC, whereas just 1% of the system DW-GEO was required for the BHE circulation pump. Due to the fact that only sensible cooling is required for the chiller unit of the hybrid system, its electricity demand was reduced by more than 50% compared to the conventional system. The overall AHU electricity demand of the system DP-VC was lower compared to the desiccant assisted systems (DW-GEO, DW-VC), because pressure drop across the desiccant wheel was saved for this system. Considering thermal energy demands, the conventional system shows the lowest thermal energy demand for reheating supply air. The required regeneration process within the desiccant assisted systems caused an increase of thermal energy demand by a factor of 1.2 compared to the system DP-VC. In total, the differences in thermal energy demand were not significantly high caused by the moderate summer period with a certain amount of oda conditions that required reheating but no or only moderate dehumidification. Nevertheless, the increased thermal energy demand of desiccant assisted air conditioning requires a convenient and low-cost heat source regarding primary energy demand of heat supply.
During winter operation, as shown in
Figure 12b, the benefits of the investigated system (DW-GEO) are not as obvious as in summer mode. Thus, electrical and thermal energy demands have to be analyzed carefully. For the reference system with electric isothermal air humidification (IH-GEO) electricity demand for air treatment and distribution was increased by a factor of 1.8 compared to DW-GEO. This was significantly induced by the electrical steam humidifier that accounts for nearly 50% of the corresponding gray column. The corresponding electricity demand for AH-GEO was reduced by 16% compared to DW-GEO for the following reasons. First, the additional electricity demand for the impeller humidifier is low and second, the pressure drop associated with the enthalpy wheel is saved. This applies for IH-GEO, respectively. Due to the reasons mentioned above, the electricity demand to operate the GCHP system is almost equal for the systems DW-GEO and AH-GEO. For the reason of high temperature steam used to humidify supply air within the system IH-GEO, required GCHP power was lower in terms of adjusting sup stream to the desired sup temperature. In total, the electricity demand of DW-GEO was reduced by 18% compared to IH-GEO and it was increased by 7% compared to AH-GEO.
Thermal energy demand required for underfloor heating is assumed to be equal for the three systems. Thermal energy required for reheating sup to the desired sup temperature was about 44% higher for the DW-GEO and AH-GEO systems compared to the system with isothermal air humidification due to substitution of thermal energy by electricity as described above. Even though the differences are quite small, considering both, electrical and thermal energy demands, system comparison shows lowest total energy demand for AH-GEO. This results from the slight advantage in electricity demand of AH-GEO against the investigated system. Thus, energetic benefits of the investigated system were limited for the considered heating period. It is expected that increased benefits will be achieved for winter terms with lower oda temperature and water content. Nevertheless, moisture recovery using an enthalpy wheel is beneficial against the reference technologies from a hygienic point of view, especially for the use of LiCl desiccant material.
Summarizing, taking summer and winter operation into account, the investigated system boasts significant reductions in electricity demand, resulting from the electricity savings during summer operation. Additional thermal energy demand required for regeneration of the desiccant wheel is not significantly higher at the same time. The required temperature level up to 70 °C can commonly be easily provided by solar thermal energy. Nevertheless, providing thermal energy as favorable as possible is crucial considering the economic efficiency of the system. Another advantage of the investigated system is the reduced amount of mode specific equipment. Subsystems like the desiccant wheel and the geothermal system of the investigated system are used throughout the year, whereas this holds not true for the chiller and the air humidifier of the reference systems.