Carnot battery integration in a geothermal power plant: study case of Zorlu Enerji

See discussions, stats, and author profiles for this publication at: https://www.researchgate.net/publication/394306376 CARNOT BATTERY INTEGRATION IN A GEOTHERMAL POWER PLANT: STUDY CASE OF ZORLU ENERJI Preprint · December 2024 CITATIONS 0 READS 26 6 authors, including: Aitor Cendoya University of Liège 26 PUBLICATIONS32 CITATIONS SEE PROFILE Chaudoir Basile University of Liège 5 PUBLICATIONS0 CITATIONS SEE PROFILE Titouan Janod University of Liège 1 PUBLICATION0 CITATIONS SEE PROFILE Tuğrul Hazar Zorlu Enerji 3 PUBLICATIONS11 CITATIONS SEE PROFILE All content following this page was uploaded by Chaudoir Basile on 05 August 2025. The user has requested enhancement of the downloaded file. 1st Belgian Symposium of Thermodynamics 16-18 December 2024, Liège, Belgium Paper 63, Page 1 CARNOT BATTERY INTEGRATION IN A GEOTHERMAL POWER PLANT: STUDY CASE OF ZORLU ENERJI Aitor Cendoya 1*, Basile Chaudoir 1*, Titouan Janod 1*, Tuğrul Hazar 2, Ural Halaçoğlu 2, Vincent Lemort 1 1 University of Liège, Thermodynamics Laboratory, Liège, Belgium 2 Zorlu Enerji, Istanbul, Türkiye * Aitor Cendoya: [email protected] * Basile Chaudoir: [email protected] * Titouan Janod: [email protected] Abstract. Waste heat generated by the industrial sector is a major source of unused energy. Integrating technologies into this process improves energy efficiency and contributes to the reduction of greenhouse gas emissions. In this scope, the SEHRENE project presents a novel approach to industrial waste heat recovery, increasing process flexibility and efficiency. The concept integrates an Electro-Thermal Energy Storage (ETES) system, including a High-Temperature Heat Pump (HTHP), a Phase Change Material (PCM) storage and an Organic Rankine Cycle (ORC). This study evaluates the optimal integration of an ETES in the largest Geothermal Power Plant (GPP) in Türkiye. The 80 MWe Kızıldere-2 GPP belongs to the Zorlu Enerji group. This plant is composed of a triple flash-binary (ORC) combined GPP. Several ETES architectures and refrigerants are analysed, to select the optimal configuration. This paper presents the results of this study case including system performance and a sensitivity analysis of various system parameters. These were then chosen to get the best design out of the trade-off between the potential performance and industrial feasibility of the machine. The main findings indicate that cyclopentane is the optimal working fluid for the given operating temperatures, where with different system architectures the round-trip efficiency (RTE) can reach values higher than 80%. Also, the performance sensitivity of the varied parameters was assessed. The most impactful varied parameters were the PCM melting temperature and the turbine isentropic efficiency. These results show the promising integration of ETES in geothermal power plants and all other processes with usable low and mediumtemperature waste heat. Keywords. High Temperature Heat Pump, Organic Rankine Cycle, Thermal Integration, Geothermal energy, ETES Nomenclature A Ċ E F h  Q  t T U v V  Heat transfer area (m²) Heat capacity rate (W/K) Energy (J) LMTD correction factor (-) Specific enthalpy (J/kg), Heat Transfer Coefficient (W/m²·K) Mass flow rate (kg/s) Heat duty (kW) Time (h) Temperature (°C or K) Overall heat transfer coefficient (W/m²·K) Specific volume (m³/kg) Volumetric flow rate (m³/s) Special characters η 𝛱𝑟 𝛥 Acronyms COP ETES GWP (HT)HP IWH KPI LMTD ORC PCM RTE VHC Efficiency (-) Pressure ratio (-) Difference (-) Coefficient Of Performance Electro-Thermal Energy Storage Global Warming Potential (High Temperature) Heat Pump Industrial Waste Heat Key Performance Indicators Log Mean Temperature Difference (K) Organic Rankine Cycle Phase Change Material Round-Trip Efficiency Volumetric Heat Capacity (J/K·m²) Paper 63, Page 2 Subscripts cd cp ev ex HX pp s sf Condenser Compressor Evaporator Exhaust Heat Exchanger Pump, Pinch point Isentropic Secondary fluid su sto tu wf x1 Supply Storage Turbine Working fluid Vapor quality of 1 1 Introduction Research on Electro-Thermal Energy Storage (ETES) has gained increasing interest because of the recent need for electricity storage to damp renewable energies’ intermittent production [1]. This specific storage solution depends less on geographical conditions [2] and on rare earth elements [3] and thus could be a suitable long-term solution. However, the complexity of these machines and the recency of the research interest are barriers to overcome. The SERHENE project aims to contribute to this field by presenting and deeply assessing novel thermally integrated designs. These designs contain High Temperature Heat Pumps (HTHP), Organic Rankine Cycles (ORC) and novel Phase Change Material (PCM) heat storages. They are also studied as to be integrated into real test cases, the one focused on here is the largest Geothermal Power Plant (GPP) in Türkiye from Zorlu Enerji. The Kızıldere-2 (KZD II) GPP has an installed capacity of 80 MWe and is composed of a triple flash-binary (ORC) combined GPP. This paper will be led as the following: • Introduction of the case study and of the reference topology. • Presentation of the thermodynamic model. • Selection of the refrigerant and the topology. • For the final topology, refinement of the parameters thanks to a sensitivity study. • Pre-sizing of the components of the system. 2 Reference Case Definition 2.1 Case study description and reference topology The reference Carnot battery design is depicted on Figure 1. Its HTHP contains an internal heat exchanger and its ORC includes a recuperator. Figure 1: ETES reference configuration. According to the information provided by Zorlu Enerji regarding the Kilzideri II three-flash plant with a Binary cycle, various energetic vectors can be recovered or utilized for potential integration into the ETES system. Table 1 presents the properties of the different vectors that could be integrated into the ETES. Based on this data, the low-pressure separator brine point will be considered as the Industrial Waste Heat (IWH) for the HTHP evaporator. Using the fluid coming from KZD II wells would have been on option but it also would have decreased the energy produced by the current system. Meanwhile, the cooling water supply to the binary cycle will be used as the cold source for the ORC condenser. Table 1: Available residual heat sources from KZDII GPP. Point T [°C] P [bar] 𝒎󰇗[ton/h] Fluid From KZD II wells (1) 178.1 10.65 2678.5 97% steam/brine Low-pressure separator brine (2) 113.1 1.59 3088.6 100% steam/brine Condensate from binary Plant (3) 50.6 3 283.9 99.9% steam/brine Cooling water supply to the binary plant (4) 24.1 3.63 9000 100% steam/brine 2.2 Thermodynamic Modelling Various components are employed and modelled, in this first stage as a thermodynamic model based on different assumptions. These hypotheses are represented for each component involved in the possible configuration of each case study. The modelling of individual components is conducted using EES 2024 (Engineering Equation Solver). 2.2.1 Heat exchangers with phase change (Evaporator and Condenser): In these components the heat transfer process involves two fluids or mediums. In the evaporator, the refrigerant or working fluid (wf) undergoes a phase change from liquid (or two-phase) to saturated, or even superheated vapour, or even to superheated vapour. Conversely, in the condenser, the refrigerant transitions from vapour (or two-phase) to liquid Paper 63, Page 3 through condensation. The energy balance between the refrigerant and the secondary fluid (sf) is presented in Equation 1. This relation corresponds to the condenser, nevertheless, for the evaporator it is the same in the reverse way. 𝑚󰇗𝑤𝑓,𝑐𝑑 ⋅(ℎ𝑤𝑓,𝑒𝑥,𝑐𝑑 −ℎ𝑤𝑓,𝑠𝑢,𝑐𝑑)= 𝑄󰇗𝑐𝑑, 𝑄󰇗𝑐𝑑 =𝑚󰇗𝑠𝑓,𝑐𝑑 ⋅(ℎ𝑠𝑓,𝑠𝑢,𝑐𝑑 −ℎ𝑠𝑓,𝑒𝑥,𝑐𝑑) (1) The saturation temperature is computed as the mean temperature between the one at the beginning and the one at the end of the two-phase zone. The equations of the pinch-point (𝑃𝑃𝑇𝐷) for the HP condenser are presented in Equations 2 and 3. These equations allow to determine by iteration of the pinch-point in the heat exchanger and set to determine the saturation state. 𝑄󰇗𝑝𝑝,𝑐𝑑 =𝑚󰇗𝑠𝑓 ⋅(ℎ𝑠𝑓,𝑝𝑝,𝑐𝑑 −ℎ𝑠𝑓,𝑠𝑢,𝑐𝑑), 𝑄󰇗𝑝𝑝,𝑐𝑑 =𝑚󰇗𝑤𝑓 ⋅(ℎ𝑤𝑓,𝑥1,𝑐𝑑 −ℎ𝑤𝑓,𝑒𝑥,𝑐𝑑) (2) Δ𝑇𝑝𝑝,𝑐𝑑 =𝑇𝑤𝑓,𝑥1,𝑐𝑑 −𝑇𝑠𝑓,𝑝𝑝,𝑐𝑑 (3) 2.2.2 Sensible heat exchangers (recuperator, preheater and superheater): These heat exchanger models only transfer sensible heat between two fluids. The recuperator transfers heat from the turbine exhaust to the pump exhaust (ORC), and after the condenser with the exhaust of the evaporator (HTHP). The preheater preheats the fluid upstream of the ORC evaporator. The superheater serves to increase the gaseous temperature at the exhaust of the ORC evaporator. The assumption for modelling these components corresponds to considering an isobaric process, then no pressure drops. A counterflow heat transfer process is considered. Finally, energy and mass are conserved in both fluids. The same energy balances are considered as in the phase change heat exchanger, adding the heat transfer equation, which is presented in Equation (4). 𝑄󰇗ℎ𝑥 =𝐶󰇗𝑚𝑖𝑛,ℎ𝑥 ⋅𝜂𝐻𝑋 ⋅(𝑇ℎ𝑜𝑡,𝑠𝑢,𝐻𝑋 −𝑇𝑐𝑜𝑙𝑑,𝑠𝑢,𝐻𝑋) (4) 2.2.3 Compressor: This component is employed in the HTHP to raise the pressure from the evaporator to the condenser. To avoid any kind of problem, superheating of the compressor suction due to refrigerant droplets must be ensured. The isentropic and volumetric efficiencies are presented in Table 2. The first is used to compute the machine real work. The second is used to compute the swept volume of the volumetric machines. The energy balance for computing the input energy for the compressor is presented in Equation 5, whereas the calculation of the swept volume is presented in Equation 6. 𝑊󰇗𝑐𝑝 =𝑚󰇗𝑤𝑓 ⋅(ℎ𝑤𝑓,𝑠,𝑒𝑥,𝑐𝑝 −ℎ𝑤𝑓,𝑠𝑢,𝑐𝑝)/𝜂𝑠,𝑐𝑝 (5) 𝑉𝑐𝑝 =𝑚󰇗𝑤𝑓 𝜂𝑣,𝑐𝑝 ·𝑁𝑐𝑝 ·𝜌𝑤𝑓 (6) 2.2.4 Turbine: This component is the machine used in the ORC to convert the fluid internal energy into electricity. This component and model can be considered identical to the compressor but with the energy transfer taking place in the opposite direction. The energy balance to compute the energy taken from fluid in the turbine corresponds to Equation 6. 𝑊󰇗𝑡𝑢 =𝑚󰇗𝑤𝑓 ⋅𝜂𝑠,𝑡𝑢 ⋅(ℎ𝑤𝑓,𝑠𝑢,𝑡𝑢 −ℎ𝑤𝑓,𝑠,𝑒𝑥,𝑡𝑢) (7) 2.2.5 Pump: The pump is the drive component in the ORC, which increases the pressure in the working fluid from the lower (condenser) to the upper pressure (evaporator). This component also consumes energy and to avoid any cavitation problems the working fluid is subcooled. The assumption for modelling these components corresponds to an isentropic process driving the fluid from the supply to the exhaust, where an efficiency of 65% was considered. The power required for the pump is computed according to Equation 7. 𝑊󰇗𝑝𝑝 = 𝑚󰇗 𝑤𝑓 ⋅(𝑣𝑤𝑓,𝑠𝑢,𝑝𝑝 ⋅(𝑃𝑤𝑓,𝑒𝑥,𝑝𝑝 −𝑃𝑤𝑓,𝑠𝑢,𝑝𝑝) 𝜂𝑠,𝑝𝑝 ) (8) 2.2.6 Phase Change Storage: PCM heat storage is the key component for the ETES which stores the energy from the HTHP and releases it in the ORC for producing electricity. The material has not been yet defined. The project’s Key Performance Indicators (KPIs) require a PCM storage with an energy density corresponding to 90𝑘𝑊ℎ/𝑚3. It is assumed that PCM can release a constant energy flow rate corresponding to the total energy storage divided by the discharging time. The energy storage in the PCM is computed according to Equation 8. 𝐸𝑠𝑡𝑜, 𝑃𝐶𝑀 = ∫ 𝑄󰇗𝑐𝑑, 𝐻𝑃 𝑑𝑡 𝑡𝑓𝑖𝑛𝑎𝑙 𝑡𝑖𝑛𝑖𝑡𝑖𝑎𝑙 (9) 2.2.7 Expansion valve: This component considers an isenthalpic process which connects the HTHP condenser with the HTHP evaporator. It is considered that this component can maintain a constant evaporator superheating. Paper 63, Page 4 Table 2 summarises the pinch point and efficiencies mentioned in the equations 3 to 8. Table 2: Assumed parameters for the reference case Parameter 𝑽𝒂𝒍𝒖𝒆 Unit Δ𝑇𝑝𝑝,𝑐𝑑 , Δ𝑇𝑝𝑝,𝑒𝑣 3 K 𝜂𝐻𝑋 0.8 - 𝜂𝑠,𝑐𝑝, 𝜂𝑠,𝑡𝑢, 𝜂𝑠,𝑝𝑝 0.65 - 𝜂𝑣,𝑐𝑝, 𝜂𝑣,𝑝𝑝 0.95 - 𝑇𝑃𝐶𝑀 160 °C 𝑡𝑐ℎ𝑎𝑟𝑔𝑒 4 h 𝑡𝑑𝑖𝑠𝑐ℎ𝑎𝑟𝑔𝑒 2 h 𝑉𝑃𝐶𝑀 322.22 𝑚3 2.3 Refrigerant Selection This section focuses on selecting refrigerants suitable for use in the ETES system. A PCM temperature of 160°C is considered in this evaluation. For HTHPs, refrigerants such as R1233zd(E), R1234(Z), R600, R601, R1336mzz(Z) and R1224yd(Z) are considered promising candidates for this application [4]. Additionally, R1234ze(Z) and R1234ze(E) are highlighted as viable options for high-temperature applications [5, 6], along with hydrocarbon-based refrigerants [7]. A comprehensive study by Frate et al. [8] has assessed 27 potential refrigerants with a Global Warming Potential (GWP) lower than 1000, identifying options that achieve critical temperatures above 125°C and maintain a saturation pressure above 0.05 bar at 40°C. Furthermore, zeotropic refrigerant mixtures have been thermodynamically evaluated, demonstrating performance improvements of up to 46% [9]. The working fluids selected for potential integration in the ETES system include Acetone, Benzene, Cyclopentane, Cyclohexane, Ethanol, Isopentane, Isohexane, Methanol, n-pentane, R1233zd(E), R1234ze(Z), R245fa, R600, and Toluene. These fluids are evaluated based on two key performance metrics for HTHP: the Coefficient of Performance (COP) and the Volumetric Heat Capacity (VHC) [9]. In ORC, the fluids are assessed based on the cycle efficiency ( 𝜂𝑜𝑟𝑐 ). Additionally, for ETES, the Round-Trip Efficiency (RTE) is also considered, calculated as the product of the HTHP COP and the ORC cycle efficiency. At first glance, refrigerants like R1234ze(Z), R245fa, and R600 are not suitable for operation at temperatures above 150°C due to limitations imposed by their critical pressures. Similarly, R1233zd(E) has a critical pressure near this threshold, which can cause operational challenges when PCM temperatures exceed 160°C. Additionally, fluids such as ethanol, methanol, cyclohexane, and toluene exhibit pressures below 0.1 bar at 20°C, which extremely increases the risk of air leakage into the ORC condenser. These refrigerants are assessed in the configuration presented in Figure 1, which integrates the reference configuration for the HTHP [4] and ORC. The main simulation results are summarized in Table 3, where refrigerants are evaluated at a HTHP heat capacity of 7 MWth, and times of charge and discharge of 4 and 2 hours, respectively. The compressor and turbine are considered at a fixed speed of 50 Hz. Table 3: Results from the simulation of the reference configuration of the pre-selected refrigerants Fluid 𝑪𝟎𝑷 𝜼𝒐𝒓𝒄 𝑹𝑻𝑬 𝑽𝑯𝑪 Acetone 4.27 0.156 0.667 3.625 Benzene 4.47 0.165 0.739 1.92 Cyclopentane 4.28 0.163 0.699 3.57 IsoPentane 3.83 0.156 0.597 4.81 IsoHexane 4.173 0.163 0.682 2.63 n-pentane 3.93 0.158 0.620 4.29 Among the fluids tested, benzene achieves the highest RTE, exceeding 70%. However, benzene’s low VHC necessitates larger components, resulting in higher investment costs compared to isopentane and n-pentane, which exhibit VHC values above 4. Acetone and cyclopentane emerge as optimal fluids, offering a balanced trade-off between RTE and VHC. Notably, cyclopentane achieves a 3.2% higher RTE and a 1.5% lower VHC compared to acetone. In conclusion, with the considered PCM temperature and the evaluated parameters, cyclopentane is determined to be the most suitable choice for the ETES system. 2.4 Architecture Selection Since the HTHP and the ORC will never work at the same time, it is possible to imagine architectures where the IWH is used during the ETES discharge. The different possible configurations analysed correspond to: • HP/ORC: Reference case (Figure 1). • HP/ORC + PRE: Reference case with integration of preheater fed with IWH. • HP/ORC + SUP: Reference case with integration of IWH to feed the evaporator and where the PCM HX is a superheater. Paper 63, Page 5 • HP/ORC + 2 stages: Reference case with integration of a 2-stages ORC. The highpressure turbine is supplied by the PCM HX and the low-pressure turbine by IWH. • ORC standalone: Only ORC, directly in contact with the IWH. For the architecture selection, a PCM temperature of 150°C is considered while all the other parameters remain the same as in Table 2. The results of the different assessed configurations are presented in Table 4. The better RTE corresponds to configuration 1, which can produce 2.168 MWe during the discharging time. Another interesting option is configuration 3, which has lower RTE performance and higher electricity output. An important remark on this configuration corresponds to the required size of the ORC is 18 and 14 times higher than cases 1 and 2, respectively. Table 4: Simulation results architecture selection. Variables HP/ ORC HP/ORC + PRE HP/ORC + SUP HP/ORC + 2stages ORC standalone 𝜂𝑜𝑟𝑐 15.9 15.7 10.9 12.5 12.2 𝐶𝑂𝑃 5.5 5.5 5.5 5.5 - 𝑅𝑇𝐸 0.87 0.86 0.60 0.68 - 𝑊󰇗𝑐𝑝 [𝑀𝑊] 1.23 1.23 1.23 1.23 - 𝑊󰇗𝑡𝑢 [𝑀𝑊] 2.17 2.62 8.36 2.14 2 The best trade-off between the RTE and output power corresponds to configuration HP/ORC + PRE, which has a RTE lower of 1.15%, but the power output increases by 20.8%. For this reason, configuration HP/ORC + PRE is highlighted as the better configuration with the given pre-design conditions. In configuration HP/ORC + SUP, the ORC efficiency is the lowest due to the fact that the ORC working fluid mass flow rate is very high in order to satisfy the discharge time as the specific heat transfer is lower. This high mass flow rate results in a drop of secondary fluid temperature in the evaporator, lowering the saturation temperature. This decreases the pressure ratio, and thus the efficiency of the ORC. The representation of the chosen architecture is presented in Figure 2. This is the selected pre-design for the Zorlu use case, this will be further analysed in the parametric section. The idea is to get the optimal values for this configuration. Figure 2: Zorlu configuration HP-ORC(Preheater) 3 Reference Case Refinement Different constraints shall be added to give a more realistic view of the ETES performance potential. These constraints and the machine parameters shall then be varied to determine which ones are the most performance sensitive and then chosen/optimized in consequence. As shown in Figure 2, the heat pump and the ORC can be studied separately except for the integrated PCM heat exchanger. Note that the impact of each parameter variation was evaluated by only considering one variation at a time. 3.1 Constraints Addition Many constraints are to be added, which are: • Constraints on heat exchanger pinch point temperature differences (PPTD). • Heat exchanger pressure drops. • Secondary fluids and their auxiliaries. First, varying heat exchanger PPTD has an impact on their working fluid saturation temperatures. This in turn impacts the saturation pressures, leading to higher/lower pressure ratios of the machine. Increasing PPTDs of HP heat exchangers increases its pressure ratio for a given condenser power, leading to lower COP. Increasing PPTDs of ORC heat exchangers decreases its pressure ratio for a given evaporator power, leading to lower efficiency. For example, the novel PCM-integrated heat exchanger still has to undergo experiments to determine its performance, a PPTD of 10K is therefore chosen as a first safety value. Secondly, heat exchanger pressure drops lower the performance of Carnot Batteries. However, some are more important than others. For the heat pump, pressure drops in the high-pressure part show little impact on its performance as the valve does not recover any work. However, the low-pressure side pressure drops tend to increase the heat pump pressure ratio, leading to higher compressor work lowering the COP. The conclusions are similar for the ORC. The pressure drops in the low-pressure side lead to a higher pressure at the outlet of the turbine, Paper 63, Page 6 as the condensing pressure is constrained to be higher than a threshold value of 0.5 bar to prevent leakages. This decreases the pressure ratio of the machine, leading to lower work production and lower ORC efficiency. The high-pressure HX pressure drops still remove useful work but this impact is lower compared to the phenomenon explained above. Depending on their importance, the allowable pressure drops assumed in heat exchangers are listed in Table 5. Table 5: HX assumed pressure drops [kPa]. Importance Liquid Gas Evaporating Condensing Low 50 20 50 50 High 30 10 15 15 Finally, secondary fluid auxiliary consumptions need to be considered since they decrease the ETES performance as they impact the heat pump COP and the ORC efficiency as follows: 𝜂𝑂𝑅𝐶 =𝑊󰇗𝑒𝑥𝑝,𝑜𝑟𝑐 − 𝑊󰇗𝑝𝑝,𝑜𝑟𝑐 −𝑊󰇗𝑝𝑝,𝑠𝑓 𝑄󰇗𝑒𝑣,𝑜𝑟𝑐 (10) 𝐶𝑂𝑃𝐻𝑃 =𝑄󰇗𝑐𝑑,𝐻𝑃 𝑊󰇗𝑐𝑝,𝐻𝑃 +𝑊󰇗𝑝𝑝,𝑠𝑓 (11) However, this negative impact can be reduced through an optimization of the secondary fluid flowrate. The optimization can be determined in this case through the tradeoff between pump consumption and the efficiency of the heat exchanger. 3.2 Heat pump parameter choice Considering the HP, the first parameters to set are the allowable heat exchanger pressure drops. These are set according to values in Table 4 and the discussion above it. Another parameter to vary is the PCM heat exchanger PPTD. From discussions with manufacturers, it was decided to fix it first to a value of 10K instead of the 3K one that was first considered. However, note that this specific heat exchanger shall be shared between the two machines. This implies that the areas computed in Section 4.1 for these two virtually independent heat exchangers (heat pump condenser and ORC evaporator) shall be similar. The biggest area is then selected because it is the most restrictive and the other PPTD is then adjusted to fit a similar area value. The PPTD for the heat pump condenser is thus adjusted from 10 K to 5 K. In addition, the heat pump secondary fluid at the evaporator has been considered. From a sensitivity study on the COP with respect to its flowrate, a value of 500 kg/s (previously 858 kg/s) was determined as optimal and set. Finally, the last parameter to consider is the PCM temperature. Lowering the PCM temperature allows the heat pump COP to increase but the ORC efficiency to decrease. In this case, the impact on the HP COP is dominant on the global system performance. Decreasing the PCM temperature increases the RTE. It was decided to vary it from 150 (in section 2.4) to 141°C. This is the lowest fusion temperature among the PCM selected for this study. Figure 3 shows the effect of the parameter variations on the heat pump COP. It can be concluded that the PCM melting temperature change has by far the most impact on performance. This is to be expected as it has a direct impact on the pressure ratio, leading to a change of compressor work and COP. The second most important factor is the increase of the PPTD at the integrated-PCM heat exchanger as it counteracts the lowering of the PCM temperature. Thirdly, the secondary fluid pump work shows a non-negligible impact on performance as it englobes the flowrate and the pressure drop effects. Finally, as already mentioned in Section 3.1, the impact of heat exchanger pressure drops highly depends on the working fluid and the pressure level that the heat exchanger is working on. The low-pressure side is the most sensitive. Figure 3: Relative effect of parameter variation on Heat Pump Performance. In conclusion, the change in parameters had a positive cumulated impact on heat pump performance thanks to the lowering of the PCM temperature. The COP of the HP improved by 0.4 (+7.3% from the previous value). Paper 63, Page 7 3.3 ORC parameter choice For the ORC, the same hypotheses as for the heat pump were considered for allowable heat exchanger pressure drops. Concerning the PCM heat exchanger PPTD, the value of 10K mentioned by manufacturers was considered. In addition, there are two secondary fluid related heat exchangers for this ORC architecture. For the sake of completeness, both of them shall be considered. The preheater secondary fluid flowrate and related pressure drop have been respectively to 60 kg/s and 50 kPa. Concerning the condenser secondary fluid, it was chosen to set its flowrate to 900 kg/s and its pressure drop to 30 kPa. These values result from a sensitivity study on ORC performance. The PCM melting temperature shall be identical to the heat pump’s one. The positive impact on the heat pump COP is higher than the negative one on the ORC efficiency when lowering the PCM temperature. Even though the ORC efficiency will suffer from it, the PCM temperature is set to 141°C. Other parameters have to be considered specifically for the ORC. First, the ORC condenser PPTD shall be varied to maintain the pressure before the pump above 50 kPa. Its value was set to 5K. Secondly, as active charge management is meant to be studied in the project, the ORC condenser subcooling was changed from 3K to 1K to get closer to a system with a liquid receiver. Finally, the ORC features a 2MWe turbomachine. The value of 65% for its isentropic efficiency seems underestimated. It was then decided to fix it to 80%, which is closer to what can be seen in the industry while not being too optimistic. Figure 4 shows the effect of the parameter variations on the ORC efficiency. It can be concluded that the most impactful change was to increase the turbine isentropic efficiency as it increases the net generated work significantly for the same turbine inlet conditions. In the ORC case, the second most impactful varied parameter type is the refrigerant heat exchanger pressure drops (again the lowpressure side ones) as they impact directly the pressure ratio of the turbine, reducing its output work. The PCM temperatures and heat exchanger PPTD had similar impacts on the ORC performance. These parameters reduce the cycle temperature difference, lowering the pressure ratio and thus, the efficiency. The difference in relative importance is mainly due to the change in Kelvins. Finally, the impact of secondary fluids is lower in the case of the ORC (especially for the preheater since the flowrate is very low). Figure 4: Relative effect of parameter variation on ORC Performance. In conclusion, the change in parameters had a negative cumulated impact on ORC performance because of low-pressure side heat exchanger pressure drop and the lowering of PCM melting temperature. The ORC efficiency dropped by 1.04 [%] (-6.72% relative to the previous value). 3.4 Global system performance To sum up, many sources of losses were added to the ETES to refine the study case. Their effects were first analysed through a sensitivity analysis together with some previously set parameters. Then, values were set (or re-set) for all these parameters. Most added parameters have a negative impact on ETES performance as they are sources of irreversibility that were not considered previously. To keep a similar RTE, the PCM temperature was decreased, and the turbine isentropic efficiency was increased. The performance results after this study are shown in Table 6. Figure 5 shows the ETES architecture with all fluid thermodynamic states. Table 6: Results for HP/ORC + PRE architecture before (B) and after (A) study case refinement. 𝜼𝒐𝒓𝒄 [%] 𝑪𝑶𝑷 [-] 𝑹𝑻𝑬 [%] 𝑾󰇗𝒄𝒑 [𝑴𝑾] 𝑾󰇗𝒏𝒆𝒕 [𝑴𝑾] 𝑻𝑷𝑪𝑴 [°C] B 15.7 5.5 86 1.23 2.62 150 A 14.4 5.9 85 1.19 2.51 141 1st Belgian Symposium of Thermodynamics 16-18 December 2024, Liège, Belgium Paper 63, Page 8 Figure 5: Chosen ETES architecture with thermodynamic states after parametric analysis. 4 Pre-Size of Components This section aims to provide a pre-design of the main ETES components. • For the heat exchangers: The technology and the heat transfer area. These are the pre-size parameters. • For the turbomachines: The work, the inlet volumetric flowrate and the pressure ratio. These are the pre-size requirements. 4.1 Heat Exchangers The heat exchanger area is pre-sized using the LMTD (Logarithmic Mean Temperature Difference) method. This method is commonly used for design and is based on temperatures at the inlet and outlet ports of a heat exchanger. It assumes that the heat duty (𝑄󰇗) can be linked to the LMTD as follows: 𝑄󰇗=𝑈𝐻𝑋 ⋅ 𝐴⋅ 𝐿𝑀𝑇𝐷⋅𝐹 (11) 𝑈𝐻𝑋 =1 1/ℎ𝑐𝑜𝑙𝑑+1/ℎℎ𝑜𝑡 (12) Where A [m²] is the heat exchanger area (to be determined), 𝑈𝐻𝑋 [W/(m² x K)] is the global heat transfer coefficient and F is the LMTD correction factor depending on the heat exchanger technology/configuration (computed using EES prebuilt dedicated functions). 𝑈𝐻𝑋 is computed by considering the convective effects of both fluid sizes, meaning that the conduction resistance in the heat exchanger wall is neglected. The heat exchanger technologies can be chosen based on their inlet fluid states and their heat duty. Due to the presence of mainly liquid secondary fluids, all heat exchangers (except the ones with PCM) were chosen to be shell-and-tube heat exchangers. The internal heat exchanger and the recuperator could feature finned tubes. From Kakaç et. al. [11], generic heat transfer coefficient values were assumed for cyclopentane (considered between a light and a medium organic fluid) and water in shell-and-tube heat exchangers. Their values are summed up in Table 7. Table 7: Typical Film Heat Transfer Coefficients for Shell-and-Tube Heat Exchangers. Heat Transfer Mode Fluid or Material 𝐔 [𝑊 𝑚²×𝐾] Sensible Liquid Water 5000 Latent solidification PCM 1000 Latent Fusion PCM 1000 Sensible Liquid Cyclopentane 1500 Sensible Gas Cyclopentane 600 Latent Evaporation Cyclopentane 2000 Latent Condensation Cyclopentane 2000 For the specific case of PCM-related heat exchanger, an equivalent heat transfer coefficient ( 𝑈𝑒𝑞,𝑃𝐶𝑀) guess proved to be challenging. Because of the discussion on the pinch point value for one mode or the other in Section 3.2, the impact of the choice of this value on the areas of both modes had to be investigated. A sensitivity analysis on the ratio of required area for both modes (HTHP condenser and ORC evaporator) with respect to this parameter was performed. The results show that the value of 𝑈𝑒𝑞,𝑃𝐶𝑀 does not impact the area ratio significantly. Indeed, when varying the parameter from 500 to 5000 W/(m² × K), the required area of the heat pump condenser remained between 55 and 60% of the area