Effects of Intercooling and Inter-Stage Heat Recovery on the Performance

energies

Article Eﬀects of Intercooling and Inter-Stage Heat Recovery on the Performance of Two-Stage Transcritical CO2 Cycles for Residential Heating Applications

Juanli Ma *, Ahmad Mhanna, Neil Juan , Monica Brands and Alan S. Fung Department of Mechanical and Industrial Engineering, Ryerson University, 350 Victoria Street, Toronto, ON M5B 2K3, Canada; [email protected] (A.M.); [email protected] (N.J.); [email protected] (M.B.); [email protected] (A.S.F.) * Correspondence: [email protected]; Tel.: +1-416-979-5000

Received: 11 November 2019; Accepted: 12 December 2019; Published: 13 December 2019

Abstract: Due to the harmful effects of synthetic refrigerants, such as chlorofluorocarbons (CFCs) and hydrochlorofluorocarbons (HCFCs)n the environment, natural refrigerants like carbon dioxide (CO2) have been attracting great interest. The higher inter-stage superheating of CO2 makes it difficult to predict the effects of the intercooling on heating performance of a two-stage transcritical CO2 cycle. In addition, very little is known about the potential of inter-stage heat rejection recovery in the heating performance enhancement of this cycle. In order to explore the effects of intercooling and inter-stage heat rejection recovery potential, three “sub-cycles”—(1) a sub-cycle with heat recovery, (2) a sub-cycle without heat recovery, and (3) a sub-cycle without intercooling—were modeled in Engineering Equation Solver (EES) software for three commonly-used two-stage transcritical cycles: (1) an intercooler cycle, (2) a flash cycle, and (3) a split cycle. Then, the discharge pressure and intermediate pressure were simultaneously optimized. Based on the optimization results, the heating performance of the sub-cycles for each cycle were compared. The results demonstrate that the incorporation of intercooling without heat recovery was detrimental to the heating performance in comparison to the absence of intercooling. It is also clear that there is a great potential for heating performance improvement through inter-stage heat recovery.

Keywords: transcritical CO2 cycle; two-stage; performance; heating applications

1. Introduction Increased concern for the world energy crisis and environmental problems has led to greater interest in eco-friendly refrigerants that are suitable for high-efficiency heat pump applications. Due to the abolition of chlorofluorocarbons (CFCs) and hydrochlorofluorocarbons (HCFCs) in the Montreal Protocol and the regulation of Hydrofluorocarbon (HFCs) in the Kyoto Protocol, the “natural” refrigerants have attracted considerable attention. Among the existing natural refrigerants, carbon dioxide (CO2) is considered especially attractive because it is nonflammable, nontoxic, free from mutagens and carcinogens, and very low in cost [1]. However, the major disadvantage of the CO2 cycle that influences its acceptance in the market is its lower performance [2–4]. Multi-stage heat pump/refrigeration cycles are typically used for large temperature differences between the source and sink, which cannot be overcome with single-stage systems [5]. Also, multi-stage cycles are an effective solution to provide power savings and improve the system performance. Two-stage cycles are usually used in food refrigeration and air conditioning for cooling applications and space or domestic water heating for heating applications, especially in cold climates. Due to the low critical temperature of CO2, in most areas of application, cycles are operated in transcritical conditions, and the existence of an “optimum” discharge pressure has received significant

Energies 2019, 12, 4763; doi:10.3390/en12244763 www.mdpi.com/journal/energies Energies 2019, 12, 4763 2 of 15

attention in the research community. In terms of the two-stage transcritical CO2 cycle, both the intermediate pressure and discharge pressure influence system performance, with each having an optimum value. Due to this unusual nature, many researchers have conducted studies on the optimization of the discharge pressure and intermediate pressure for two-stage transcritical CO2 cycles. In cooling applications, Groll and Kim [6] conducted a literature review of the available research on transcritical CO2 cooling cycles. They reported that the ideal intermediate pressure has frequently been found to differ from the classical estimation of the geometric mean of the evaporator and discharge pressure. The formula for the classical estimation is √Pev Pd, where Pev is the pressure in the × evaporator and Pd is the discharge pressure of the second-stage compressor. Additionally, an optimum discharge pressure has generally been found to exist, with a higher gas cooler temperature leading to a higher optimum discharge pressure. Hwang et al. [7] measured the experimental performance of intercooler and split cycles, and found that every cycle displayed an optimum discharge pressure. Cavallini et al. [8] performed a theoretical analysis on a split cycle and an experimental analysis on an intercooler cycle. They investigated the effect of the second-stage pressure ratio on performance. Manole [9] found that the optimum intermediate pressure differed from the classical estimate also. Agrawal et al. [10] simultaneously optimized the discharge pressure and inter-stage pressure for flash and intercooler cycles, and found an optimum intermediate pressure and optimum discharge pressure for each cycle. Ozgur [11] performed a theoretical simulation of an intercooler cycle and determined the optimum discharge pressure for various gas cooler outlet temperatures. Ozgur and Bayrakci [12] studied the effect of the intermediate pressure on performance and concluded that there was an optimum intermediate pressure that maximized both first-law and second-law efficiencies. Cecchinato et al. [13] studied the split, flash, and intercooler systems experimentally, and confirmed that the optimum intermediate pressure deviated from the classical estimate by examining different intermediate pressures at specific discharge pressures. Srinivasan [14] studied an intercooler system and also found that the optimal intermediate pressure differed from the classical estimate. Almeida and Barbosa [15] simulated transcritical CO2 cycles with and without intercooling. They found that intercooling provided a significant increase in the coefficient of performance (COP). Ozgur and Tosun [16] simulated flash and intercooler cycles and noted that the intermediate pressure had a greater effect on flash cycles. Zhang et al. [17] examined the effect of the discharge pressure, intermediate pressure, and mass flow rate on various flash and intercooler systems, and found an optimum value for all parameters investigated. Bush et al. [3] tested a lab-scale two-stage system to explore the effects of the mechanical subcooling on the system performance. A steady-state model for the system was developed and presented. To improve the performance of the transcritical CO2 system, different combinations of a liquid suction heat exchanger after-cooler and two-stage compression were embedded into the system configuration by Mohammadi [4]. The modified configurations were modeled in detail using Engineering Equation Solver (EES) software, and energy and exergy analyses were performed for each configuration. Due to the large throttle loss in the expansion valve, the two-stage cycles with advanced expansion devices, such as ejector, expander, and vortex tubes, have been of interest in the research community in recent years. Manjili and Yavari [18] proposed a new two-stage multi-intercooling refrigeration cycle employing an ejector, where the performance of this cycle was compared with two one-stage ejector refrigeration cycles. Sun et al. [19] compared the performance of six transcritical CO2 cycles with and without an expander-compressor, which used an expander as an expansion device and served as an assistant compressor or the main compressor. Several different expander-compressor arrangements, including the two-stage cycle, were investigated. Bayrakci et al. [20] also analyzed the expander usage in a two-stage transcritical CO2 cooling system. Variable parameters in this study were the gas cooler pressure, inter-stage pressure, and evaporation temperature of the refrigerant. Xing et al. [21] proposed a cycle with two ejectors as expansion devices. The performance of the improved two-stage cycles were evaluated and then compared with those of the basic two-stage cycle with a flash tank. Nemati et al. [22] compared the performance of a two-stage ejector-expansion transcritical refrigeration cycle Energies 2019, 12, 4763 3 of 15

using ethane and CO2 as refrigerants. The theoretical analysis of the cycle performance characteristics was carried out for both refrigerants according to the first and second laws of thermodynamics. Zhang et al. [23] conducted a sensitivity study for three typical expander-based transcritical CO2 cycles, including a two-stage cycle. The sensitivities of the maximum COP to the key operating parameters, including the inlet pressure of the gas cooler, the temperatures at evaporator inlet and gas cooler outlet, the inter-stage pressure, and the isentropic efficiency of expander, were obtained. In heating applications, Wang et al. [24] tested the heating COP variation with discharge pressure and intermediate pressure. The result confirmed both of the pressures’ influence on the overall system heating performance, which is similar to that in cooling performance. Pitarch et al. [25] analyzed and optimized a two-stage split CO2 cycle for heating applications regarding COP in terms of the discharge pressure. The classical estimate of the intermediate pressure was employed in this analysis. This finding differs from cooling cycle in that the heating COP decreases as the refrigerant is cooled down at the intercooler. Many researchers have dealt with the two-stage transcritical CO2 system for cooling applications while few paid attention to cycles for heating applications. In fact, the inter-stage superheating of compressed CO2 vapor is far higher than that of traditional refrigerants [13]. For example, when R134a, R410a, and CO2 are compressed between an evaporating temperature of 270 K and a condensing temperature of 290 K, R134a is superheated by only 2.88 K and R410a is superheated by 8.89 K, while CO2 is superheated by 16.57 K. This feature leads to the difficulty in predicting whether the intercooling has a beneficial effect on the heating performance of the two-stage cycle. This is because the removal of heat decreases the second-stage work but it also reduces the amount of heat that can be released from the gas cooler. As a result, it is not apparent which of these opposing effects has a greater influence on heating COP. In addition, it would be a waste of energy if the inter-stage heat is rejected to the ambient environment. However, very little is known about the potential of inter-stage heat rejection recovery in the heating performance enhancement of two-stage transcritical CO2 cycles. In order to understand the effects of intercooling and inter-stage heat rejection recovery on the performance of two-stage transcritical CO2 cycles for heating applications, three “sub-cycles” have been optimized and compared based on the models developed in EES 10.0 software (University of Wisconsin, WI, USA) for two-stage transcritical CO2 cycles. Because both the two pressures could affect the performance, unlike using classical estimate of intermediate pressure in Pitarch et al. [25], the two optimal pressures for each sub-cycle were identified simultaneously in this study. The three sub-cycles were: (1) a sub-cycle with heat recovery (with HR), (2) a sub-cycle without heat recovery (without HR), and (3) a sub-cycle without intercooling (without IC). In terms of two-stage cycle selection for investigation, due to the unavailability of advanced expansion technology in the market and the findings that the throttling valve heating cycles are able to be applied to the advanced expansion cycles, in this study, the three commonly-used cycles with a throttling valve were investigated: (1) a basic two-stage compression intercooler cycle, (2) a flash cycle, and (3) a split cycle. This paper is organized as follows. First, the optimum discharge pressure from the second-stage compressor (PD) and intermediate pressure (PM) for each sub-cycle were identified simultaneously. Second, the heating coefficients of performance (COPs) of the sub-cycles over a range of evaporating temperatures were compared. Finally, the cause of the observed trends in the COP for each sub-cycle . was explained based on heating capacity (Q) and compression work (W).

2. Cycles and Sub-Cycles under Analysis Figures1–3 present the schematics and P–h diagrams of the intercooler cycle, the flash cycle, and the split cycle. Here are the descriptions of the three commonly-used cycles investigated: (1) Intercooling cycle: The intercooler cycle incorporates a low-pressure (LP) and high-pressure (HP) compressor and a single expansion valve. The two compressors work in series and interact with the mass flow leaving the evaporator. (2) Flash Cycle: The cycle operates similarly to the basic intercooling cycle, but with the addition of a flash tank and an expansion valve. The CO2 expanding after the gas cooler enters a Energies 2019, 12, 4763 4 of 15 Energies 2019, 12, x FOR PEER REVIEW 4 of 15 Energies 2019, 12, x FOR PEER REVIEW 4 of 15 theflash HP tankcompressor at the inter-stage and the liquid pressure, will be where throttled the vapor to the will evaporator be drawn pressure. by the HP (3) compressor Split Cycle: and The the the HP compressor and the liquid will be throttled to the evaporator pressure. (3) Split Cycle: The cycleliquid incorporates will be throttled an LP to and the HP evaporator compressor pressure. in series, (3) Split and Cycle: two expansion The cycle incorporates valves. The anmass LP andflow HP cycle incorporates an LP and HP compressor in series, and two expansion valves. The mass flow splitscompressor after the ingas series, cooler, and where two expansionone part expands valves. through The mass the flow valve splits and after enters the the gas evaporator, cooler, where and one splits after the gas cooler, where one part expands through the valve and enters the evaporator, and thepart other expands part expands through and the is valve injected and with enters the the outgoing evaporator, stream and of the the other low-pressure part expands compressor. and is injected the other part expands and is injected with the outgoing stream of the low-pressure compressor. withTo theinvestigate outgoing the stream effects of theof intercooling low-pressure and compressor. inter-stage heat rejection recovery on the heating To investigate the effects of intercooling and inter-stage heat rejection recovery on the heating performanceTo investigate of the above the e cycles,ffects of three intercooling sub-cycles and we inter-stagere evaluated heat for rejection each of recoverythe selected on cycles. the heating In performance of the above cycles, three sub-cycles were evaluated for each of the selected cycles. In thisperformance analysis, the of heat the aboverejected cycles, from three the gas sub-cycles cooler was were used evaluated for space for heating each of and theselected the civil cycles. water was In this this analysis, the heat rejected from the gas cooler was used for space heating and the civil water was usedanalysis, as coolant the heat for the rejected intercooler. from the The gas heat cooler recovered was used through for space civil heating water andcan thebe used civil waterfor domestic was used used as coolant for the intercooler. The heat recovered through civil water can be used for domestic hotas water coolant (e.g., for preheating the intercooler. or direct The utilizatio heat recoveredn), which through makes civilthis sub-cycle water can more be used practical. for domestic hot hot water (e.g., preheating or direct utilization), which makes this sub-cycle more practical. water (e.g., preheating or direct utilization), which makes this sub-cycle more practical.

(a) (b) (a) (b) Figure 1. Intercooler cycle: (a) schematic and (b) P–h diagram. HP: high pressure, LP: low pressure. Figure 1. Intercooler cycle: (a) schematic and (b) P–h diagram. HP: high pressure, LP: low pressure. (TheFigure numbers 1. Intercooler (1, 2, 3, 4, cycle: 5, 6) in (a ()a schematic) refer to the and position (b) P–h of diagram. the working HP: highfluid; pressure, The numbers LP: low (1, 2, pressure. 3, 4, (The numbers (1, 2, 3, 4, 5, 6) in (a) refer to the position of the working fluid; The numbers (1, 2, 3, 4, 5, 6)(The in ( numbersb) refer to (1, the 2, 3,thermodynamic 4, 5, 6) in (a) refer state to point the position of the working of the working fluid in ﬂuid; P–h diagram.). The numbers (1, 2, 3, 4, 5, 5, 6)6) in in (b (b) )refer refer to to the the thermodynamic thermodynamic state state point point of of the the working working fluid ﬂuid in in P–h P–h diagram.). diagram.).

(a) (b) (a) (b) FigureFigure 2: Flash 2. Flash cycle: cycle: (a) (schematica) schematic and and (b) (P–hb) P–h diagram. diagram. (The (The numbers numbers (1, (1,2, 3, 2, 4, 3, 5, 4, 6) 5, in 6) ( ina) (refera) refer to to Figure 2: Flash cycle: (a) schematic and (b) P–h diagram. (The numbers (1, 2, 3, 4, 5, 6) in (a) refer to thethe position position of the of the working working fluid; ﬂuid; The The numbers numbers (1, (1,2, 3, 2, 4, 3, 5, 4, 6) 5, in 6) ( inb) ( brefer) refer to the to the thermodynamic thermodynamic state state the position of the working fluid; The numbers (1, 2, 3, 4, 5, 6) in (b) refer to the thermodynamic state pointpoint of ofthe the working working fluid ﬂuid in inP–h P–h diagram.). diagram.). point of the working fluid in P–h diagram.).

EnergiesEnergies 20192019, ,1212, ,x 4763 FOR PEER REVIEW 5 5of of 15 15

(a) (b)

FigureFigure 3: 3. SplitSplit cycle: cycle: ( (aa)) schematic schematic and and (b) P–hP–h diagram.diagram. (The(The numbersnumbers (1, (1, 2, 2, 3, 3, 4, 4, 5, 5, 6) 6) in in (a )(a refer) refer to theto theposition position of theof the working working ﬂuid; fluid; The The numbers numbers (1,2, (1, 3, 2, 4, 3, 5, 4, 6) 5, in 6) (b in) refer(b) refer to the to thermodynamicthe thermodynamic state state point pointof the of working the working ﬂuid fluid in P–h in diagram.).P–h diagram.).

TheThe layout layout of the sub-cyclessub-cycles areare as as follows: follows: (1) (1) Sub-cycle Sub-cycle with with heat heat recovery: recovery: In theIn sub-cyclethe sub-cycle with withheat heat recovery, recovery, an intercooler an intercooler was installedwas installed after afte ther first-stage the first-stage compressor. compressor. Both Both the heat the heat rejected rejected from fromthe gas the cooler gas cooler and the and heat the rejected heat rejected from the from intercooler the intercooler were included were forincluded heating for capacity heating calculations. capacity calculations.(2) Sub-cycle (2) without Sub-cycle intercooling: without intercooling: In the sub-cycle In the without sub-cycle intercooling, without intercooling, no intercooler no was intercooler installed wasafter installed the first-stage after the compressor. first-stage (3)compressor. Sub-cycle (3) without Sub-cycle heat without recovery: heat In therecovery: sub-cycle In the without sub-cycle heat withoutrecovery, heat an intercoolerrecovery, an was intercooler installed afterwas installe the first-staged after compressor.the first-stage However, compressor. only theHowever, heat rejected only thefrom heat the rejected gas cooler from was the included gas cooler for was heating included capacity for calculations.heating capacity Table calculations.1 shows the Table features 1 shows of the thethree features sub-cycles. of the three sub-cycles

TableTable 1. 1. TheThe features features of of the the three three sub-cycles. sub-cycles.

NameName of the Sub-CycleSub-Cycle Feature Feature Abbreviation Abbreviation Sub-cycleSub-cycle with heatheat recovery recovery With With both both intercooling intercooling and and heat heat recovery recovery with with HR HR Sub-cycleSub-cycle without intercoolingintercooling With With neither neither intercooling intercooling orheat or heat recovery recovery without without IC IC Sub-cycleSub-cycle without heatheat recovery recovery With With only only intercooling intercooling without without HR HR

The heating COP of each sub-cycle for the intercooler cycle was defined as follows: The heating COP of each sub-cycle for the intercooler cycle was defined as follows: 𝐶𝑂𝑃 = = , (1) . . . Q m (h h)+m((h h ) 𝐶𝑂𝑃 =with HR =2 2 3 4 4 , 5 (2) COPwith HR = . = . − .− , (1) ( ) + (( ) Wwith HR m1 h2 h1 m3 h4 h3 𝐶𝑂𝑃 = = − −. (3) . The heating COP of each sub-cycle for the flash cycle was. defined as follows: Qwithout IC m4 ((h4 h5) COP = = 0 0 − , (2) 𝐶𝑂𝑃without IC = . = . . , W m1(h2 h1) + m2 ((h4 h2) (4) without IC − 0 − . . 𝐶𝑂𝑃 = = , (5) Qwithout HR m4(( h4h5) COP = . = . −. . (3) without HR ( )+ (( ) 𝐶𝑂𝑃 W= =m1 h2 h1 m3 h4. h3 (6) without HR − − TheThe heating heating COP COP of of each each sub-cycle sub-cycle for for the the split flash cycle cycle was was defined defined as as follows: follows: 𝐶𝑂𝑃 = = , (7) . . . Q m (h h ) + m (h h ) with HR 2 2 34 4 5 COP𝐶𝑂𝑃with HR = =. = =. − − , , (8) (4) ( )+ (( ) Wwith HR m1 h2 h1 m3 h4 h3 − − 𝐶𝑂𝑃 = = , (9) . . Q m ((h h ) where 𝑄 is the heating capacity (kJ/s); without𝑊 is ICthe compression40 work40 (kJ/s);5 𝑚 is the mass flowrate COPwithout IC = = . . − , (5) of CO2 (kg/s); ℎ is the specific enthalpyW withoutof CO2 IC (kJ/kg).m1( h2 h1) + m3 ((h4 h3 ) − 0 0 − 0

Energies 2019, 12, 4763 6 of 15

. . Qwithout HR m4((h4 h5) COPwithout HR = . = . −. . (6) W m1(h2 h1) + m3((h4 h3) without HR − − The heating COP of each sub-cycle for the split cycle was deﬁned as follows:

. . . Qwith HR m2(h2 h3) + m4((h4 h5) COPwith HR = . = . − . − , (7) W m1(h2 h1) + m3((h4 h3) with HR − − . . (( ) Qwithout IC m40 h40 h5 COPwithout IC = . = . . − , (8) W m1(h2 h1) + m3 ((h4 h3 ) without IC − 0 0 − 0 . . Qwithout HR m4((h4 h5) COPwithout HR = . = . − , (9) W m1(h2 h1) + m3((h4 h3) without HR − − . . . where Q is the heating capacity (kJ/s); W is the compression work (kJ/s); m1 is the mass ﬂowrate of CO2 (kg/s); h is the speciﬁc enthalpy of CO2 (kJ/kg).

3. Mathematical Modelling

3.1. Thermodynamic Analysis An Engineering Equation Solver (EES) [26] is a numerical equation-solving program with many built-in mathematical and thermophysical property functions useful for engineering calculations. The high accuracy thermodynamic and transport property database means that EES is widely used in thermodynamic and heat transfer fields. The reliability and the accuracy of the numerical models in this study depend on the compressor efficiency correlation, the temperature calculation of the internal heat exchanger (IHX), and the temperature difference between CO2 and water, which are presented as follows from the former research results. The efficiency of the compressor is estimated by employing the following correlation for the semi-hermetic compressor [27]:

2 P P η = 0.26 + 0.7952 c,o 0.2803 c,o − Pc,i − Pc,i 3 4 (10) P P +0.0414 C,O 0.0022 c,o , Pc,i − PC,i where η is the compressor eﬃciency, Pc,o is the CO2 pressure at the outlet of the compressor (MPa), and Pc,i is the CO2 pressure at the inlet of the compressor (MPa). For the split cycle, the temperatures for the high-pressure IHX could be given by [25]:

T6 = T9 + 3. (11)

The outlet enthalpy of the expansion valve could be calculated using:

houtlet = hinlet, (12) where houtlet is the CO2 enthalpy at the outlet of the expansion valve (kJ/kg) and hinlet is the CO2 enthalpy at the inlet of the expansion valve (kJ/kg). Llopis et al. [28] tested a CO2 transcritical refrigeration plant, where the difference between the gas cooler outlet CO2 temperature and inlet water temperature was reported to be less than 5 ◦C. Hence, the temperature difference between the inlet water and outlet CO2 was taken as 5 ◦C in this study. In practical applications, the superheat is controlled at a constant to adjust the refrigerant mass flow rate to meet the various demands. In simulated analyses, the constant is usually selected to be a small value, like 0 ◦C, 5 ◦C, or 10 ◦C. The superheat assumption will affect the optimization results, but Energies 2019, 12, 4763 7 of 15 a similar trend will be found for different superheat assumptions. In this study, the superheat at the inlet of the first stage compressor was assumed to be 0 ◦C. Compared to the high operating pressure of the CO2 cycle, the pressure drop in the pipes and heat exchanger is very small. Hence, the pressure drop in the pipes and heat exchangers were considered to be negligible. This assumption has little influence on the thermodynamic state parameters and thus the optimization results.

3.2. Optimization Conditions In this paper, each cycle was optimized regarding the maximum heating COP using the conjugate directions method in EES. The discharge pressure and the intermediate pressure were simultaneously optimized. Because the critical temperature and critical pressure of CO2 are 31.1 ◦C and 7.39 MPa, this means the heat is rejected in a supercritical process in residential heating applications. Hence, the lower bound of the PD for the second stage compressor was taken to be 7.4 MPa and the upper bound of the PM, namely the discharge pressure of the first stage compressor, was taken to be 6.8 MPa. The manufacturer could provide a CO2 compressor capable of operating at the maximum of 14 MPa; therefore, the upper bound of the PD was taken to be 14 MPa. For the two sub-cycles with heat recovery and the sub-cycle without heat recovery, civil water was used as a coolant for the intercooler. If the temperature in the intercooler is lower than that of civil water, the civil water neither has an effect on cooling down the intercooler nor has the capability of recovering the heat rejected from the intercooler. With the assumption of the 5 ◦C temperature difference between the civil water temperature and the CO2 temperature (discussed in Section 3.1), the lower bound of PM for these two sub-cycles was set to be 5.1 MPa, where 5.1 MPa is the corresponding saturated pressure of CO2 at 15 ◦C. Table2 presents the optimization conditions.

Table 2. The optimization conditions of the three sub-cycles.

Name of the Sub-Cycles PD PM Lower bound Upper bound Lower bound Upper bound Sub-cycle with heat recovery 7.4 MPa 14 MPa 5.1 MPa 6.8 MPa Sub-cycle without intercooling 7.4 MPa 14 MPa 2.5 MPa 6.8 MPa Sub-cycle without heat recovery 7.4 MPa 14 MPa 5.1 MPa 6.8 MPa

4. Results and Discussion The cycle performance was evaluated on the basis of a maximum heating COP to obtain optimum values for PD and PM. The performance was evaluated over an evaporator temperature range of 20 C to 0 C. The heat rejected from gas cooler was used for space heating, so the gas cooler outlet − ◦ ◦ temperature of CO2 was ﬁxed at 40 ◦C to meet the requirements of space heating. This temperature could be achieved by adjusting the inlet water ﬂow rate according to the heat demand. The performance parameters and their optimum values are displayed graphically and elucidated below.

4.1. Optimization Results

4.1.1. Intercooler Cycle

Figure4 shows the variation of the optimum PD and PM for each sub-cycle with the change of Tev from 20 C to 0 C for the intercooler cycle. It was found that for the sub-cycle without heat recovery, − ◦ ◦ the optimum PM was calculated to have an optimum value of 5.1 MPa for all operating conditions. This means the that sub-cycle without heat recovery had a maximum COP when PM was operating at its minimum limit. The sub-cycle without intercooling had an optimum PM of 5.1 MPa when the Tev was less than 15 C. It was also observed that the optimum P of each sub-cycle decreased linearly − ◦ D with Tev. For the sub-cycle with heat recovery, the optimal PD decreased linearly with Tev, while PM increased with Tev. Energies 2019, 12, x FOR PEER REVIEW 7 of 15 the lower bound of PM for these two sub-cycles was set to be 5.1 MPa, where 5.1 MPa is the corresponding saturated pressure of CO2 at 15 °C. Table 2 presents the optimization conditions.

Table 2. The optimization conditions of the three sub-cycles.

4. Results and Discussion The cycle performance was evaluated on the basis of a maximum heating COP to obtain optimum values for PD and PM. The performance was evaluated over an evaporator temperature range of −20 °C to 0 °C. The heat rejected from gas cooler was used for space heating, so the gas cooler outlet temperature of CO2 was fixed at 40 °C to meet the requirements of space heating. This temperature could be achieved by adjusting the inlet water flow rate according to the heat demand. The performance parameters and their optimum values are displayed graphically and elucidated below.

4.1. Optimization Results

4.1.1. Intercooler Cycle

Figure 4 shows the variation of the optimum PD and PM for each sub-cycle with the change of Tev from −20 °C to 0 °C for the intercooler cycle. It was found that for the sub-cycle without heat recovery, the optimum PM was calculated to have an optimum value of 5.1 MPa for all operating conditions. This means the that sub-cycle without heat recovery had a maximum COP when PM was operating at its minimum limit. The sub-cycle without intercooling had an optimum PM of 5.1 MPa when the Tev was less than −15 °C. It was also observed that the optimum PD of each sub-cycle decreased linearly withEnergies Tev2019. For, 12 ,the 4763 sub-cycle with heat recovery, the optimal PD decreased linearly with Tev, while8 of P 15M increased with Tev.

FigureFigure 4. 4. OptimumOptimum discharge discharge pressure pressure and and intermedia intermediatete pressure pressure for the intercooler cycle. 4.1.2. Flash Cycle 4.1.2. Flash Cycle Figure5 shows the variation of the optimum PD and PM for each sub-cycle with the change of Figure 5 shows the variation of the optimum PD and PM for each sub-cycle with the change of Tev Tev from 20 ◦C to 0 ◦C for the ﬂash cycle. With the similarity of the intercooler cycle, the sub-cycle from −20 −°C to 0 °C for the flash cycle. With the similarity of the intercooler cycle, the sub-cycle withoutEnergies 2019 heat, 12,recovery x FOR PEER had REVIEW an optimum PM of 5.1 MPa. The sub-cycle without intercooling had8 of an15 without heat recovery had an optimum PM of 5.1 MPa. The sub-cycle without intercooling had an optimum PM of 5.1 MPa at Tev equals 20 ◦C. It was also observed that the optimum PD decreased optimum PM of 5.1 MPa at Tev equals −−20 °C. It was also observed that the optimum PD decreased almost linearly withwith TTevev for all the sub-cyclessub-cycles whilewhile PPMM increased with Tevev forfor the sub-cycle with heat recovery. The optimum PD ofof the sub-cycle without HR showed a signiﬁcantsignificant decrease, while the optimum PD had a slight decrease for the other two sub-cycles.

Figure 5. Optimum discharge pressure and intermediate pressure for ﬂash cycle. Figure 5. Optimum discharge pressure and intermediate pressure for flash cycle. 4.1.3. Split Cycle 4.1.3. Split Cycle Figure6 shows the variation of the optimum PD and PM for each sub-cycle with the change of Figure 6 shows the variation of the optimum PD and PM for each sub-cycle with the change of Tev Tev from 20 C to 0 C for the split cycle. It can be observed that, regarding the optimal P of the − ◦ ◦ M sub-cyclefrom −20 °C without to 0 °C heat for recovery,the split cycle. the split It can cycle be observed resembled that, the regarding other two the cycles optimal investigated PM of the above. sub- cycle without heat recovery, the split cycle resembled the other two cycles investigated above. The The sub-cycle without heat recovery had an optimum PM of 5.1 MPa until Tev was around 5 ◦C. sub-cycle without heat recovery had an optimum PM of 5.1 MPa until Tev was around −5 °C. Like− the Like the ﬂash cycle, the optimum PD decreased almost linearly with the increase of the Tev for all the flash cycle, the optimum PD decreased almost linearly with the increase of the Tev for all the sub-cycles sub-cycles and the PM increased with the increase of the Tev for the sub-cycle with heat recovery and theand sub-cyclethe PM increased without with intercooling. the increase This of might the T beev for because the sub-cycle the sub-cycle with heat without recovery heat recoveryand the sub- had cycle without intercooling. This might be because the sub-cycle without heat recovery had the the maximum COP at the lower bound of the PM when the Tev was less than 5 ◦C. Therefore, the PD maximum COP at the lower bound of the PM when the Tev was less than −−5 °C. Therefore, the PD without HR was larger than the PD with HR when the Tev was less than 5 ◦C and lower than the PD without HR was larger than the PD with HR when the Tev was less than −−5 °C and lower than the PD with HR when Tev was 0 ◦C and 5 ◦C. with HR when Tev was 0 °C and −5 °C.

Figure 6. Optimum discharge pressure and intermediate pressure for split cycle.

From the optimal PD and PM values of the three cycle types, it can be concluded that the optimal PD decreased with Tev and the optimal PM increased with Tev for two-stage heating operations. These findings agree with the results of Agrawal et al. [10] for cooling operations.

4.2. Effects on Performance

4.2.1. Intercooler Cycle

Energies 2019, 12, x FOR PEER REVIEW 8 of 15 almost linearly with Tev for all the sub-cycles while PM increased with Tev for the sub-cycle with heat recovery. The optimum PD of the sub-cycle without HR showed a significant decrease, while the optimum PD had a slight decrease for the other two sub-cycles.

Figure 5. Optimum discharge pressure and intermediate pressure for flash cycle.

4.1.3. Split Cycle

Figure 6 shows the variation of the optimum PD and PM for each sub-cycle with the change of Tev from −20 °C to 0 °C for the split cycle. It can be observed that, regarding the optimal PM of the sub- cycle without heat recovery, the split cycle resembled the other two cycles investigated above. The sub-cycle without heat recovery had an optimum PM of 5.1 MPa until Tev was around −5 °C. Like the flash cycle, the optimum PD decreased almost linearly with the increase of the Tev for all the sub-cycles and the PM increased with the increase of the Tev for the sub-cycle with heat recovery and the sub- cycle without intercooling. This might be because the sub-cycle without heat recovery had the maximum COP at the lower bound of the PM when the Tev was less than −5 °C. Therefore, the PD without HR was larger than the PD with HR when the Tev was less than −5 °C and lower than the PD Energies 2019, 12, 4763 9 of 15 with HR when Tev was 0 °C and −5 °C.

FigureFigure 6. 6. OptimumOptimum discharge discharge pressure pressure and intermediate pressure for split cycle. Energies 2019, 12, x FOR PEER REVIEW 9 of 15 From the optimal P and P values of the three cycle types, it can be concluded that the optimal From the optimal PD and PMM values of the three cycle types, it can be concluded that the optimal P decreasedFigure with 7 Tillustratesand the the optimal comparisonP increased of the withCOP Tforfor all two-stagesub-cycles heating operating operations. at optimum PD decreased with Tev evand the optimal PM increasedM with Tev for evtwo-stage heating operations. These conditions for the intercooler cycle. It can be seen that the sub-cycle with heat recovery had the findingsThese ﬁndings agree with agree the with results the results of Agrawal of Agrawal et al. [10] et al. for [10 cooling] for cooling operations. operations. highest COP and the sub-cycle without heat recovery had the lowest COP. As Tev varied from −20 °C 4.2. Eﬀtoects 0 °C, on Performancethe sub-cycle with heat recovery experienced an increase in COP from 2.4 to 3.3. The COP of 4.2. Effects on Performance the sub-cycle with heat recovery was greater than that of the sub-cycle without intercooling by 11.2% 4.2.1. Intercooler Cycle 4.2.1. Intercoolerto 14.1%, and Cycle greater than that of the sub-cycle without heat recovery by 24.2% to 50.3%. FigureThe7 illustrates results show the comparison that the recovery of the COP of heat for from all sub-cycles the intercooler operating yielded at optimum the highest conditions COP among for thethe intercooler sub-cycles. cycle. The ItCOP can of be the seen sub-cycle that the with sub-cycle heat recovery with heat was recovery even grea hadter the than highest the COP COP of the and thesub-cycle sub-cycle without without intercooling, heat recovery with had an theadvantage lowest COP.of about As T13.1%.ev varied It had from a 37.5%20 Cadvantage to 0 C, the over the − ◦ ◦ sub-cyclesub-cycle with heat without recovery heat experienced recovery. However, an increase installing in COP an from intercooler 2.4 to 3.3. without The COP heat of the recovery sub-cycle after the with heatfirst-stage recovery compression was greater led than to a that COP of about the sub-cycle 17.4% lower without than intercooling that of the sub-cycle by 11.2% towith 14.1%, no intercooler and greaterat than all. that of the sub-cycle without heat recovery by 24.2% to 50.3%.

Figure 7. Heating coefficient of performance (COP) for the intercooler cycle at an optimum discharge Figure 7. Heating coefficient of performance (COP) for the intercooler cycle at an optimum discharge and intermediate pressure. and intermediate pressure. The results show that the recovery of heat from the intercooler yielded the highest COP among The COP trends for the intercooler cycle is explained via the differing compression work and the sub-cycles. The COP of the sub-cycle with heat recovery was even greater than the COP of the heating capacity of the sub-cycles. Figure 8 illustrates the heating capacity and compression work per sub-cycle without intercooling, with an advantage of about 13.1%. It had a 37.5% advantage over the unit of CO2 mass for the sub-cycles operating at optimum conditions in the intercooler cycle. sub-cycle without heat recovery. However, installing an intercooler without heat recovery after the first-stage compression led to a COP about 17.4% lower than that of the sub-cycle with no intercooler at all. The COP trends for the intercooler cycle is explained via the differing compression work and heating capacity of the sub-cycles. Figure8 illustrates the heating capacity and compression work per unit of CO2 mass for the sub-cycles operating at optimum conditions in the intercooler cycle.

Figure 8. Heating capacity and compression work for intercooler cycle at an optimum discharge and intermediate pressure.

It can be observed that both the heating capacity and compression work of the sub-cycle without intercooling were greater than those of the sub-cycle with heat recovery. Although both the heating capacity and work values of the sub-cycle without intercooler were greater than those of the sub- cycle with heat recovery, the higher COP of the sub-cycle with heat recovery was attributed to the compression work having a more significant influence on COP than heating capacity. For example, when Tev = −20 °C, the heating capacity of the sub-cycle without intercooling was greater than that of

Energies 2019, 12, x FOR PEER REVIEW 9 of 15

Figure 7 illustrates the comparison of the COP for all sub-cycles operating at optimum conditions for the intercooler cycle. It can be seen that the sub-cycle with heat recovery had the highest COP and the sub-cycle without heat recovery had the lowest COP. As Tev varied from −20 °C to 0 °C, the sub-cycle with heat recovery experienced an increase in COP from 2.4 to 3.3. The COP of the sub-cycle with heat recovery was greater than that of the sub-cycle without intercooling by 11.2% to 14.1%, and greater than that of the sub-cycle without heat recovery by 24.2% to 50.3%. The results show that the recovery of heat from the intercooler yielded the highest COP among the sub-cycles. The COP of the sub-cycle with heat recovery was even greater than the COP of the sub-cycle without intercooling, with an advantage of about 13.1%. It had a 37.5% advantage over the sub-cycle without heat recovery. However, installing an intercooler without heat recovery after the first-stage compression led to a COP about 17.4% lower than that of the sub-cycle with no intercooler at all.

Figure 7. Heating coefficient of performance (COP) for the intercooler cycle at an optimum discharge and intermediate pressure.

The COP trends for the intercooler cycle is explained via the differing compression work and heatingEnergies 2019 capacity, 12, 4763 of the sub-cycles. Figure 8 illustrates the heating capacity and compression work10 ofper 15 unit of CO2 mass for the sub-cycles operating at optimum conditions in the intercooler cycle.

Figure 8. Heating capacity and compression work for intercooler cycle at an optimum discharge and Figure 8. Heating capacity and compression work for intercooler cycle at an optimum discharge and intermediate pressure. intermediate pressure. It can be observed that both the heating capacity and compression work of the sub-cycle without It can be observed that both the heating capacity and compression work of the sub-cycle without intercooling were greater than those of the sub-cycle with heat recovery. Although both the heating intercooling were greater than those of the sub-cycle with heat recovery. Although both the heating capacity and work values of the sub-cycle without intercooler were greater than those of the sub-cycle capacity and work values of the sub-cycle without intercooler were greater than those of the sub- with heat recovery, the higher COP of the sub-cycle with heat recovery was attributed to the compression cycle with heat recovery, the higher COP of the sub-cycle with heat recovery was attributed to the work having a more significant influence on COP than heating capacity. For example, when Tev = 20 ◦C, compression work having a more significant influence on COP than heating capacity. For example,− the heating capacity of the sub-cycle without intercooling was greater than that of the sub-cycle with when Tev = −20 °C, the heating capacity of the sub-cycle without intercooling was greater than that of heat recovery by 5.3%. Meanwhile, the compression work of the sub-cycle without intercooling was greater than that of the sub-cycle with heat recovery by 17.3%. As a result, the sub-cycle with heat recovery outperformed the sub-cycle without intercooling. Regarding the COP comparison of the sub-cycle without intercooling and the sub-cycle without heat recovery, both the heating capacity and compression work of the sub-cycle without intercooling were greater than those of the sub-cycle without heat recovery. The difference in heating capacity ranged between 37.6% and 22.2%, and the difference in work ranged between 9.5% and 8.5% as the evaporator temperature rose. It can be concluded that the decrease in the heating capacity was proportionally greater. Therefore, the sub-cycle without heat recovery had a lower COP compared to the sub-cycle without intercooling. Regarding the COP comparison of the sub-cycle with heat recovery and the sub-cycle without heat recovery, the situation was different. The sub-cycle with heat recovery had a higher heat capacity and a lower compression work, leading to a better COP.

4.2.2. Flash Cycle Figure9 illustrates the comparison of the COP for all sub-cycles at optimum conditions for the flash cycle. The COP of the sub-cycle with heat recovery increased with Tev from 2.7 to 3.7, and was greater than that of the sub-cycle without intercooling by 9.8% to 10.4%. It was also greater than that of the sub-cycle without heat recovery by 34.3% to 19.7% as the evaporator temperature increased. In addition, the COP of the sub-cycle without heat recovery was less than that of the sub-cycle without intercooling by 17.7% to 7.8% as Tev increased. The results show that for the flash cycle, the recovery of heat from the intercooler was also beneficial to the performance. The COP of the sub-cycle with heat recovery was greater than the COP of the sub-cycle without intercooling by about 10.2%, and greater than the COP of the sub-cycle without heat recovery by about 26.0%. Additionally, installing an intercooler after the first-stage compression without heat recovery led to a lower COP than having no intercooler; the COP of the sub-cycle without heat recovery was less than that of the sub-cycle without intercooling by around 12.4%. Energies 2019, 12, x FOR PEER REVIEW 10 of 15

the sub-cycle with heat recovery by 5.3%. Meanwhile, the compression work of the sub-cycle without intercooling was greater than that of the sub-cycle with heat recovery by 17.3%. As a result, the sub- cycle with heat recovery outperformed the sub-cycle without intercooling. Regarding the COP comparison of the sub-cycle without intercooling and the sub-cycle without heat recovery, both the heating capacity and compression work of the sub-cycle without intercooling were greater than those of the sub-cycle without heat recovery. The difference in heating capacity ranged between 37.6% and 22.2%, and the difference in work ranged between 9.5% and 8.5% as the evaporator temperature rose. It can be concluded that the decrease in the heating capacity was proportionally greater. Therefore, the sub-cycle without heat recovery had a lower COP compared to the sub-cycle without intercooling. Regarding the COP comparison of the sub-cycle with heat recovery and the sub-cycle without heat recovery, the situation was different. The sub-cycle with heat recovery had a higher heat capacity and a lower compression work, leading to a better COP.

4.2.2. Flash Cycle Figure 9 illustrates the comparison of the COP for all sub-cycles at optimum conditions for the flash cycle. The COP of the sub-cycle with heat recovery increased with Tev from 2.7 to 3.7, and was greater than that of the sub-cycle without intercooling by 9.8% to 10.4%. It was also greater than that of the sub-cycle without heat recovery by 34.3% to 19.7% as the evaporator temperature increased. In addition, the COP of the sub-cycle without heat recovery was less than that of the sub-cycle without intercooling by 17.7% to 7.8% as Tev increased. The results show that for the flash cycle, the recovery of heat from the intercooler was also beneficial to the performance. The COP of the sub-cycle with heat recovery was greater than the COP of the sub-cycle without intercooling by about 10.2%, and greater than the COP of the sub-cycle without heat recovery by about 26.0%. Additionally, installing an intercooler after the first-stage compression without heat recovery led to a lower COP than having no intercooler; the COP of the

Energiessub-cycle2019, 12, 4763 without heat recovery was less than that of the sub-cycle without intercooling11 by of 15around 12.4%.

FigureFigure 9. Heating 9. Heating COP for COP the for flash the cycle flash atcycle an optimum at an opti dischargemum discharge and intermediate and intermediate pressure. pressure. Figure 10 illustrates the heating capacity and compression work values of all sub-cycles at optimum Figure 10 illustrates the heating capacity and compression work values of all sub-cycles at conditions for the flash cycle. Regarding the COP comparison of the sub-cycle with heat recovery optimum conditions for the flash cycle. Regarding the COP comparison of the sub-cycle with heat and the sub-cycle without intercooling, similar to the intercooler cycle, both the heating capacity recovery and the sub-cycle without intercooling, similar to the intercooler cycle, both the heating and compression work of the sub-cycle without intercooling were greater. The difference in heating capacity and compression work of the sub-cycle without intercooling were greater. The difference in capacity ranged between 7.9% and 7.7%, and the difference in work ranged between 19.9% and 18.9% heating capacity ranged between 7.9% and 7.7%, and the difference in work ranged between 19.9% with increasing evaporator temperature. Since the higher compression work was a more significant and 18.9% with increasing evaporator temperature. Since the higher compression work was a more factor than the higher heating capacity, the sub-cycle with heat recovery outperformed the sub-cycle significant factor than the higher heating capacity, the sub-cycle with heat recovery outperformed the withoutEnergies 2019 intercooling., 12, x FOR PEER REVIEW 11 of 15 sub-cycle without intercooling.

Figure 10. Heating capacity and compression work for the flash ﬂash cycle at an optimum discharge and intermediate pressure.

In terms of the COP comparison of the sub-cycl sub-cyclee without intercooling an andd the sub-cycle without heat recovery, the heating capacity of the sub-cyclesub-cycle without intercooling was greater than that of the sub-cycle without heat recovery byby 34.4%34.4% toto 17.9%17.9% whenwhen TTevev increased. The The compression compression work of the sub-cycle without intercoolingintercooling waswas moremore than than that that of of the the sub-cycle sub-cycle without without heat heat recovery recovery by by 11.3% 11.3% to to8.7% 8.7% as asTev Tincreased.ev increased. Obviously, Obviously, the the greater greater heating heating capacity capacity ofof thethe sub-cyclesub-cycle withoutwithout intercooling was a moremore significantsignificant factorfactor thatthat itsits betterbetter performance.performance. Hence, the sub-cycle without intercooling yielded a higher COP than the sub-cycle without heatheat recovery.recovery. Regarding the COP comparison of the sub-cycle with heat recovery and the sub-cycle without heat recovery, the compressioncompression work of the sub-cycle with heat recovery was lower and its heating capacity was higher compared to the sub-cycle withou withoutt heat recovery. Hence, the sub-cycle with heat recovery yielded a higher COP than the sub-cycle withoutwithout heatheat recovery.recovery. From thethe aboveabove discussion,discussion, it it can can be be concluded concluded that that the the reasons reasons for for the the diff differingering COPs COPs among among the theflash flash sub-cycles sub-cycles was was the samethe same as for as thefor intercoolerthe intercooler cycle. cycle.

4.2.3. Split Cycle Figure 11 illustrates the COP of all sub-cycles at optimum conditions for the split cycle. It can be seen that the COPs of this cycle had the same behavior as in the intercooler and flash cycles. The COP of the sub-cycle with heat recovery increased from 2.9 to 4.0 in the range investigated. The COP of the sub-cycle with heat recovery was greater than that of the sub-cycle without intercooling by between 10.8% and 11.4%. The COP of the sub-cycle with heat recovery was greater than that of the sub-cycle without heat recovery by 28.2% to 23.7%. The COP of the sub-cycle without heat recovery was less than that of the sub-cycle without intercooling by a value ranging from 13.6% to 9.9% with rising Tev.

Figure 11. Heating COP for split cycle at optimum discharge and intermediate pressure.

Energies 2019, 12, x FOR PEER REVIEW 11 of 15

Figure 10. Heating capacity and compression work for the flash cycle at an optimum discharge and intermediate pressure.

In terms of the COP comparison of the sub-cycle without intercooling and the sub-cycle without heat recovery, the heating capacity of the sub-cycle without intercooling was greater than that of the sub-cycle without heat recovery by 34.4% to 17.9% when Tev increased. The compression work of the sub-cycle without intercooling was more than that of the sub-cycle without heat recovery by 11.3% to 8.7% as Tev increased. Obviously, the greater heating capacity of the sub-cycle without intercooling was a more significant factor that its better performance. Hence, the sub-cycle without intercooling yielded a higher COP than the sub-cycle without heat recovery. Regarding the COP comparison of the sub-cycle with heat recovery and the sub-cycle without heat recovery, the compression work of the sub-cycle with heat recovery was lower and its heating capacity was higher compared to the sub-cycle without heat recovery. Hence, the sub-cycle with heat recovery yielded a higher COP than the sub-cycle without heat recovery. From the above discussion, it can be concluded that the reasons for the differing COPs among

Energiesthe2019 flash, 12, 4763 sub-cycles was the same as for the intercooler cycle. 12 of 15

4.2.3. Split Cycle 4.2.3. Split Cycle Figure 11 illustrates the COP of all sub-cycles at optimum conditions for the split cycle. It can be Figureseen that 11 illustratesthe COPs of the this COP cycle of allhad sub-cycles the same atbehavior optimum as conditionsin the intercooler for the and split flash cycle. cycles. It can The be COP seen thatof the the sub-cycle COPs of thiswith cycle heat had recovery the same increased behavior from as in2.9 the to intercooler4.0 in the range and ﬂash investigated. cycles. The The COP COP of of thethe sub-cycle sub-cycle with with heat heat recovery recovery increased was fromgreater 2.9 tothan 4.0 that in the of range the sub-cycle investigated. without The COPintercooling of the by sub-cyclebetween with 10.8% heat recovery and 11.4%. was The greater COP than of the that sub-cycle of the sub-cycle with heat without recovery intercooling was greater by than between that of the 10.8%sub-cycle and 11.4%. without The COP heat of recovery the sub-cycle by 28.2% with to heat 23.7%. recovery The COP was of greater the sub-cycle than that without of the sub-cycle heat recovery withoutwas heat less recovery than that by of 28.2%the sub-cycle to 23.7%. without The COP interc ofooling the sub-cycle by a value without ranging heat from recovery 13.6% was to 9.9% less with than thatrising of T theev. sub-cycle without intercooling by a value ranging from 13.6% to 9.9% with rising Tev.

Energies 2019, 12, x FOR PEER REVIEW 12 of 15 Figure 11. FigureHeating 11. Heating COP forCOP split for cycle split atcycle optimum at optimum discharge discharge and intermediate and intermediate pressure. pressure. The results results show show that that for for the the split split cycle, cycle, the the recovery recovery of of heat heat from from the the intercooler intercooler still still resulted resulted in thein the highest highest COP, COP, while while the the installation installation of of an an intercooler intercooler without without heat heat recovery recovery after after the the first-stage ﬁrst-stage compressioncompression still resulted in the the lowest lowest COP. COP. The CO COPP of the the sub-cycle sub-cycle with with heat heat recovery recovery was was greater greater thanthan the the COP COP of of the the sub-cycle sub-cycle without without intercooling intercooling by by about about 10.9%, 10.9%, and and greater greater than the COP of the sub-cyclesub-cycle without without heat heat recovery by about 25.4%. In Installationstallation of of an an intercooler intercooler after after the first-stage ﬁrst-stage compressioncompression reduced the COP compared to the sub-cyclesub-cycle without intercooling byby aroundaround 11.4%.11.4%. Figure 1212 illustrates illustrates the the heating heating capacity capacity and compressionand compression work comparison work comparison at optimal at conditions optimal conditionsfor the split for cycle. the split The resultscycle. The may results be due may to the be sub-cycle due to the without sub-cycle heat without recovery heat having recovery an optimum having an optimum PM at the lower bound until Tev was around −5 °C. Consequently, the heating capacity PM at the lower bound until Tev was around 5 ◦C. Consequently, the heating capacity slope of the slope of the sub-cycle without HR changed at− around −5 °C. This led to the heating capacity of the sub-cycle without HR changed at around 5 ◦C. This led to the heating capacity of the sub-cycle sub-cycle without HR being greater than the− sub-cycle with HR when Tev was more than −5 °C. without HR being greater than the sub-cycle with HR when Tev was more than 5 C. − ◦

Figure 12. HeatingHeating capacity capacity and and compression compression work work for for split split cycle cycle at at optimum optimum discharge discharge and intermediateintermediate pressure. pressure.

The heating capacity and and compression work work of of th thee sub-cycle without without inte intercoolingrcooling were were greater greater thanthan those those of of the the sub-cycle sub-cycle with with heat heat recovery. recovery. Th Thee difference diﬀerence in heating capacity ranged ranged from from 4.1% 4.1% to 3.7% and the difference in work ranged from 16.5% to 15.2% with rising evaporator temperature. This was due to the fact that the higher compression work appeared to be more significant than the difference in heating capacity. The sub-cycle with heat recovery had a higher COP than the sub-cycle without intercooling. Regarding the COP comparison of the sub-cycle with heat recovery and the sub-cycle without heat recovery, the higher heating capacity and lower compression work of the sub-cycle with heat recovery yielded a relatively higher COP. The sub-cycle without intercooling yielded a higher COP than the sub-cycle without heat recovery for the same reason: lower compression work and higher heating capacity. The reasons for the differing COPs among the sub-cycles of the split cycle was different from the above two cycles. Regarding the effect of intercooling on performance, the sub-cycle without intercooling yielded a higher COP than the sub-cycle without heat recovery because of its lower compression work and higher heating capacity.

5. Conclusions In this study, the effects of intercooling on the heating performance of the two stage transcritical CO2 cycle and inter-stage heat rejection recovery potential were explored for three commonly used two-stage cycles—intercooler, flash, and split cycles—through a performance comparison of three sub-cycles. When the sub-cycles operated at optimal conditions, the main findings were as follows: 1. There was a great potential for performance improvement via the recovery of heat from the intercooler. The COP of the sub-cycle with heat recovery was greater than that of the sub-cycle without intercooling by around 13.1% for the basic intercooler cycle, 10.2% for the flash cycle,

Energies 2019, 12, 4763 13 of 15 to 3.7% and the difference in work ranged from 16.5% to 15.2% with rising evaporator temperature. This was due to the fact that the higher compression work appeared to be more significant than the difference in heating capacity. The sub-cycle with heat recovery had a higher COP than the sub-cycle without intercooling. Regarding the COP comparison of the sub-cycle with heat recovery and the sub-cycle without heat recovery, the higher heating capacity and lower compression work of the sub-cycle with heat recovery yielded a relatively higher COP. The sub-cycle without intercooling yielded a higher COP than the sub-cycle without heat recovery for the same reason: lower compression work and higher heating capacity. The reasons for the differing COPs among the sub-cycles of the split cycle was different from the above two cycles. Regarding the effect of intercooling on performance, the sub-cycle without intercooling yielded a higher COP than the sub-cycle without heat recovery because of its lower compression work and higher heating capacity.

5. Conclusions In this study, the effects of intercooling on the heating performance of the two stage transcritical CO2 cycle and inter-stage heat rejection recovery potential were explored for three commonly used two-stage cycles—intercooler, flash, and split cycles—through a performance comparison of three sub-cycles. When the sub-cycles operated at optimal conditions, the main findings were as follows:

1. There was a great potential for performance improvement via the recovery of heat from the intercooler. The COP of the sub-cycle with heat recovery was greater than that of the sub-cycle without intercooling by around 13.1% for the basic intercooler cycle, 10.2% for the flash cycle, and 10.9% for the split cycle; furthermore, it was greater than that of the sub-cycle without heat recovery by around 37.5% for the basic intercooler cycle, 26.0% for the flash cycle, and 25.4% for the split cycle. 2. Incorporating an intercooler without heat recovery reduced the COP compared to the sub-cycle without intercooling for all cycle types. The COP of the sub-cycle without heat recovery was less than that of the sub-cycle without intercooling by around 17.4% for the basic intercooler cycle, 12.4% for the flash cycle, and 11.4% for the split cycle. 3. The sub-cycle with heat recovery had a better COP than the sub-cycle without intercooling for all three cycle types examined; this was because the sub-cycle with heat recovery had lower compression work and lower heating capacity. However, the lowered compression work was a more significant factor, which led to a better performance. Similarly, the lower COP of the sub-cycle without heat recovery compared to the sub-cycle without intercooling was attributed to the fact that the decrease in heating capacity was more significant than the difference in compression work.

4. It was found that the optimum discharge pressure, PD, decreased with the evaporating temperature, Tev, and the optimum intermediate pressure, PM, increased with Tev. This result was consistent with the results for cooling operations in the literature.

Author Contributions: All authors contributed to the research in this paper: conceptualization, J.M.; software, A.M., N.J., and J.M.; data curation, A.M., M.B., and J.M.; writing—original draft preparation, J.M. and M.B.; writing—review and editing, A.S.F.; visualization, J.M.; supervision, project administration, and funding acquisition, A.S.F. Funding: This research was funded by a Natural Sciences and Engineering Research Council (NSERC) of Canada Discovery Grant and the Ontario Centre of Excellence (OCE). Conﬂicts of Interest: The authors declare no conﬂict of interest. Energies 2019, 12, 4763 14 of 15

Nomenclature

Pc,o Pressure at the outlet of the compressor (MPa) Pc,i Pressure at the inlet of the compressor (MPa) PD Discharge pressure of the second-stage compressor (MPa) PM Inter-stage pressure (MPa) . W Compression work (kJ/s) . Q Heating capacity (kJ/s) η Compressor efficiency (-) . m1 Mass flowrate (kg/s) h Specific enthalpy (kJ/kg)

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