**3. Results and Discussion**

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 (950 ± 95) m<sup>3</sup> ·h−1; mass flow rates of supply and extract air were controlled to be equal. Set point of sup water content is 8 gw ·kg−<sup>1</sup> air 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 di fferent 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 Wel 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 ( *P*GCHP = 887 − 1388 Wel). 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]:

$$\text{COP}\_{\text{el,AHU,su}} = \frac{\dot{m}\_{\text{sup}} \cdot (h\_1 - h\_{4'})}{P\_{\text{el,AHU}}} \quad \text{COP}\_{\text{el,AHU,sw}} = \frac{\dot{m}\_{\text{sup}} \cdot (h\_{4'} - h\_1)}{P\_{\text{el,AHU}}} \tag{1}$$

$$\text{COP}\_{\text{th,A-HU,su}} = \frac{\dot{m}\_{\text{sup}} \cdot (h\_1 - h\_{\text{4}^\circ})}{\dot{Q}\_{\text{th,A-HU,su}}} \quad \text{COP}\_{\text{th,A-HU,su}} = \frac{\dot{m}\_{\text{sup}} \cdot (h\_1 - h\_{\text{4}^\circ})}{\dot{Q}\_{\text{th,A-HU,su}}} \tag{2}$$

. *Q*th,AHU,su = . *<sup>m</sup>*w,AH · (*h*w,in,AH − *<sup>h</sup>*w,out,AH) . *Q*th,AHU,wi = . *<sup>m</sup>*w,RH · (*h*w,in,RH − *<sup>h</sup>*w,out,RH) (3)

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 COPel,AHU,su 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, COPel,AHU,su 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 COPel,AHU,su with increasing regeneration air temperature. Due to its definition, negative values of COPel,AHU,su occur at low outside air temperatures when dehumidification of supply air is necessary whereas cooling is not required.

**Figure 4.** (**a**) Electrical COP of the air handling unit during summer operation at dehumidification mode in dependence of regeneration air temperature; (**b**) thermal COP of the air handling unit during summer operation at dehumidification mode in dependence of regeneration air temperature.

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 COPth,AHU,su at higher regeneration air temperature. This characteristic results from the mathematical definition of COPth,AHU,su as presented in Equation (2). Negative values of COPth,AHU,su 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 COPth,AHU,su ≥ 1 with decreasing fluctuations for increasing regeneration air temperature.

For the investigated heating period, performance indicators in form of COPel,AHU,wi and COPth,AHU,wi are shown in Figure 5. To evaluate the performance of the air handling unit during winter mode, the characteristics of COPel,AHU,wi and COPth,AHU,wi depending on outside air temperature are presented for EW mode and HRW mode.

**Figure 5.** (**a**) Electrical COP of the air handling unit during winter operation in dependence of outside air temperature; (**b**) thermal COP of the air handling unit during winter operation in dependence of outside air temperature.

.

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 su fficient 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 COPel,AHU,wi 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, COPth,AHU,wi is independent of outside air temperature and not dropping below COPth,AHU,wi = 2. 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 e fficiency is defined as follows: .

$$\text{DCOP} = \frac{\dot{m}\_{\text{sup}} \cdot (\mathbf{x}\_1 - \mathbf{x}\_{4'}) \cdot r\_0}{\dot{Q}\_{\text{th,AHU,su}}} \, , \tag{4}$$

with *Q*th,AHU,su according to Equation (3) and is therefore equivalent to the definition of latent COP. An average dehumidification e fficiency of DCOP = 1.15 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 e fficiency 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 e fficiency is expressed by:

$$\Psi = \frac{\mathbf{x}\_2 - \mathbf{x}\_1}{\mathbf{x}\_7 - \mathbf{x}\_1}.\tag{5}$$

An average moisture recovery e fficiency of Ψ = 0.75 was achieved for the enthalpy wheel during the investigated winter period. Thus, an increase of 1.1 g w kg−<sup>1</sup> air was achieved for sup humidity ratio on average. Maximum values of moisture recovery were close to 2.3 gw kg−<sup>1</sup> air 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 su fficient 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.

**Figure 6.** (**a**) Temperature profile of the grouting material for both borehole heat exchangers; (**b**) soil energy balance based on thermal energy input and output.

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 0.22 MWhth, 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 (*Q*BHE) and heat flow ( . *Q*BHE) 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):

$$\text{MPF} = \frac{\int\_{\text{m}} \left| \dot{Q}\_{\text{BHE}} \right| \, \text{d}\tau}{\int\_{\text{m}} P\_{\text{PU}} \, \text{d}\tau} \quad \text{SPF} = \frac{\int\_{\text{p}} \left| \dot{Q}\_{\text{BHE}} \right| \, \text{d}\tau}{\int\_{\text{p}} P\_{\text{PU}} \, \text{d}\tau} \tag{6}$$

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 MPFsu,max = 170 ± 17, 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 SPFsu = 153 ± 15 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 SPFwi = 110 ± 11, 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:

$$\text{COP}\_{\text{GCHP}} = \frac{\left| \dot{Q}\_{\text{h}} \right|}{P\_{\text{GCHP}} + P\_{\text{AUX}}} \tag{7}$$

**Figure 7.** Thermal energy transfer at the BHE (**left**) and performance parameters of the BHE (**right**) for the investigated periods: (**a**) cooling period; (**b**) heating period.

**Figure 8.** (**a**) GCHP electrical COP in dependence of BHE outlet temperature ϑBHE,out; (**b**) total GCHP temperature lift <sup>Δ</sup>ϑGCHP,tot.

BHE outlet temperatures were mostly pooled in the range of ϑBHE,out = 3.8 − 8 ◦C during steady state GCHP operation with a range of COPGCHP = 2.9 − 3.5, as shown in Figure 8a. A trend of increasing COPGCHP 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 <sup>Δ</sup>ϑGCHP,tot is defined as temperature difference between condenser outlet and evaporator inlet. Values above <sup>Δ</sup>ϑGCHP,tot = 32 K 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):

$$\text{SPF}\_{\text{GCHP}} = \frac{\int\_{\text{P}} \left| \dot{Q}\_{\text{h}} \right| \, \text{d}\tau}{\int\_{\text{P}} (P\_{\text{GCHP}} + P\_{\text{AUX}}) \, \text{d}\tau} \tag{8}$$

For the considered heating period SPFGCHP = 3 was achieved. Compared to GCHP systems state of the art with SPFGCHP = 4.0 − 4.5, performance of the investigated system relying on a reciprocating compressor shows potential for improvement. Nevertheless, this system is robust against fluctuating temperature levels.
