TY - JOUR
AU1 - Yunus, Mohammed
AU2 - Oreijah, Mowffaq M
AB - ABSTRACT Among non-toxic and inflammable working fluids, carbon dioxide (CO2) proving a possible choice of refrigerant is gaining full status because of its reducing capacity of global warming and ozone depletion. Heat pump (HP) equipment uses energy more rationally to heat water by reducing emissions and global warming caused by conventional refrigerants. HP capability demands an environmentally friendly refrigerant for their full utilization and high energy saving as well to attaining higher coefficient of performance (COP) without much design corrections. CO2 is checked for sustainability to eliminate the standardized refrigerants such as chlorofluorocarbons (CFC) and hydro-CFC. In this work, the performance of CO2 as an alternate refrigerant in HP for heating water at different pressure, flow rates, evaporator fan speed and preheating have been investigated. HP is accommodated with two condensers. The results showed that the increased mass flow rate had increased COP, overall heat transfer (HT) coefficient (U) and HT. However, logarithmic mean temperature difference was decreased for increasing evaporator’s fan speed and pressure. The outlet water temperature (TWO) of second condenser increased with decreasing water flow rate. Improved COP, HT and TWO of HP are observed from the experimental evaluation in case of preheating of water. 1. INTRODUCTION The refrigerants and their properties used in a heat pump (HP) system run based on a reversible Rankine cycle. They employed to warm up the houses or cool the systems during the winter and summer seasons to continuously provide conditioned weather. The dichloro-difluoro-methane (R-22) and tetra-fluoro-ethane (R-134a) having main constituents as chlorofluorocarbons (CFC) are widely obtainable and currently in use as refrigerants. The increasing CFC rate within the globe is heading toward ‘global warming potential (GWP)’ and ‘ozone depletion potential (ODP)’ issues. Hence, it becomes necessary to shift to new natural refrigerants contributing no ODP and minimum GWP and substitute to synthetic type like carbon dioxide (CO2) in HP. CO2 is considered because of its merits, like no toxicity, no non-combustible, inexpensive, commendable thermo-physical characteristics and risk-free surroundings. The regular refrigerant CO2 has been recognized as artificial refrigerants, causing two significant GWP and ODP threats worldwide. The key to solve these problems depends on the intensifying efficiency of energy utilization and swapping to free carbon energies [1]. Generally, most fitting applications of this refrigerant are found in domestic water heating. HP consists of four components, expansion valve, evaporator, compressor and condenser, and works by taking advantage of the refrigerant characteristics during the condensing and evaporating stages. In contrast, the compressor provides high pressure and temperature (P&T) vapor refrigerant to the condenser [2]. The heat will be absorbed by the cold water circulating in the covering surface like a counter heat flow exchange method. The refrigerant was sent to an expansion valve like a capillary tube and then sent to the evaporator at lower P&T in the form of vapor to reduce refrigerant pressure. The fan furnished in the evaporator increases the heat exchange between air and vapor refrigerant for changing it into low pressure to enter the compressor. This way, the cycle gets repeated. There are vapor compression (VC) and vapor absorption types of HP [3]. The cycle’s significant adaptations to review the output and its optimum conditions are proposed comprehensively by using multi-phasing, internal ‘heat exchanger (HE)’, divergent-type turbine, expulsor, whirlpool tube, etc. on the transcritical CO2 cycle [4]. An exhaustive comparison carried out on optimum conditions with pressure on the high side showed the maximum improvement by changing the turbine through an active recovery or multi-phasing despite being an expensive stage. Investigations generally focused on the ejector-expansion cycle proved significant because it improved COP, involves no moving parts, and inexpensive [5]. A simulation model of HP water heater utilizing the CO2 to validate the experimental results of its performances containing the gas cooler (GC), evaporator, compressor and expansion valve was proposed. The water inlet and external air temperature on the output descriptions showed a difference in COP results of the two methods [6]. The performance (COP and water outlets temperatures) evaluation on the transcritical CO2-HP using the experimentation and simulation model for real-time heating and cooling impact of water on their flowing rate, GC, etc. was investigated. The water flow rate and inlet temperature significantly affected system performances using two GC on the single internal tube. In contrast, three inner screwed tubes showed decreased boiling capacity ratio, weight and pressure drop of the waterside [7]. The low critical temperature of CO2 (31.10°C) and modified CO2 VC system in standard refrigeration (HP and air conditioning; A/C) showed temperatures worked almost or slightly higher than designated pressure (7.38 MPa) of evaporation, which is called a ‘transcritical cycle (TCC)’ [8]. The GC of HE cools CO2 gas, showing COP decreasing with increasing pressure compared to TCC, and the temperature loss during heat rejection is more than refrigerant cycles. Using counterflow HE, temperature loss reduced in heating using coolant for the single-stage ‘heat transfer (HT)’ in HPs. The heating capacity (HC) and COP using vaporization temperature affected little compared to other refrigerants that have promoted the CO2 for sustaining a higher HC even at low intake temperature [9]. Numerical simulation and experimental prototype model evaluated the performance of TCC-CO2-HP under simultaneous cooling and heating and showed HT capability of CO2 improved for both two-stage and supercritical section [10]. The results of the COP of TC- CO2 refrigerator with internally mounted HE showed improvements up to 10% over HP with GC to reduce loss during the expansion process. The increasing COP is found in a 2-phase compressor utilizing a flash intercooler performing internal cooling of vapor by vaporization except at primary working conditions [11]. Comparing experimental with traditional direct expansion method, COP, cooling ability and ‘coefficient of HT (h)’ were increased but decreased pressure drop at the refrigerant side when the flash gas bypass was employed. Among the various problems on the usage of CO2 as a refrigerant, the high pressure for operating is handled by mixing it with available refrigerants [12]. HT trials for the supercritical CO2 (SCC) flow into a 10.9 mm inner diameter tube and were compared to the experimental data and mathematical model of it [13]. Numerical study and experimental validation of turbulent SCC cooled in the tube were developed [14]. Numerical modeling of GC to compare a range of mathematical relationships of 1-stage HT proved a useful method for temperature difference model in forecasting the HT process of SCC instead of the traditional logarithmic mean temperature difference (LMTD) method. A concept of compacted HE for CO2 A/C systems uses a 2 mm inner diameter tube to reduce 70% of HT losses compared to the HFC system as a result of high flow velocity and density [15]. The compressor performance maps using test data of 2-CO2 model compressors (‘semi-hermetic’, ‘double-piston’, ‘single-phase’, ‘reciprocating’ and a ‘hermetic’, ‘2-phase’, ‘rotary type’) subjected to load variations were prepared [16]. Three micro-channel tubes piled up to offer several parallel openings to control pressure drop [17]. Maximizing h and areas upstream of the expansion have reduced 50% material requirements and increased 10% efficiency [18]. The internal HE effect on the TC-CO2 system’s performance in different applications like mobile A/C was studied theoretically [19]. The ozone layer in the earth’s climate contains the accessible particle of three oxygen iotas (O3) layer [20]. The excessive release of bromine and chlorine mixtures from a CFC, ‘hydro bromo fluoro carbons (HBFC)’, ‘methyl chloroform (CH3CCl3)’ and ‘hydro-CFC (HCFC)’ depletes O3 [21]. O3 degradation that occurs due to refrigeration activities, which release CFCs, is the main reason. The global warming effect occurs due to an increase in the troposphere’s temperature by increasing greenhouse gasses (GHGs) contents. GHGs absorb the infrared radiation of sun rays reflected by the earth and remain in the atmosphere. The present ‘average temperature of the Earth (ATE)’ is 15°C. Increasing GHG (from CO2, methane, CFC, water vapor, nitrous oxide and O3) to the environment, ATE is rising steadily, which has already climbed up to 0.5°C from the past few decades [22]. Nevertheless, the effects of flow rate, preheating of water and increasing condensers on HP performance are not studied much [23]. Few environmentally friendly refrigerants were also used but could reduce GHG only to a certain extent. A VC refrigerator working on R134a refrigerant has been tested with hydrocarbons (natural refrigerants) like liquefied petroleum gas (LPG). The results of LPG consisting of 60% propane (R290) and 40% butane (R600) showed both the target temperature and break-up time were improved. The COP improved, and the power consumption decreased compared to R134a [24]. The effect of the recovering HE characteristics on HP efficiency using new HP refrigerant by mixing with some cooling agents was investigated by experimental and numerical modeling. The results of replacing tubular (double-pipe) with a plate HE to reduce the flow and increase the compressor inlet’s heating temperature revealed improved efficiency [25]. Different low-carbon processes, schemes, regimes, engaged population preparation, progressive green structure approaches, renewable type energy supplying units, etc. to encourage low carbon living and energy-related behavior were thoroughly discussed [26]. Experimental and numerical analysis were performed for finding the optimum configuration of micro-channels with five sets of rectangular configurations under forced convection heat transfer with a constant base area. Theoretical analysis with a C program code found optimal dimension at various flow inputs and heat inputs. It was found that the HT coefficient (h) for a hydraulic diameter of 260 μm is higher as compared to the other that is more optimum [27]. Performance of micro-channel HE in residential HP A/C systems compared with that of the baseline HP system using tube-and-fin type HE showed the system capacity could increase by 4% and efficiency by 1–2%, keeping the same face area of the HE. The weight of HE and refrigerant charge was reduced by 44% and 51%, respectively [28]. Investigations on the Kilimanjaro Region's energy-consumption pattern of Tanzania for switching energies, its possible aid in reducing deforestation and CO2 emission. The study results that showed for most of the accused in energy consumption are utilizing three-stone type stoves with their favored energy resource that is firewood. If wood-saving stoves are provided, then it could reduce firewood utilization and CO2 emission by 51%, which has been suggested [29]. The VC-HP with four refrigerants was studied to find high-temperature HP’s efficiency. The COP for heating was dependent on the low-temperature source level of the condenser temperature. The recommended refrigerants are R236fa and R600a that are to be employed to heat industrial cooling water [30]. The current work is on modifying the HP with two condensers and an evaporator with a fan for testing the CO2 as an alternate refrigerant in a water heating system to meet the requirements of Saudi Arabia’s climatic conditions. Such conventional water heaters consuming more energy and harmful to the environment can be replaced with the HP using CO2 refrigerant. The performance of naturally available refrigerants like CO2 will be analyzed to minimize GWP and to maintain zero ODP. Investigations on HP’s performance like heating capacities, system COP, outlet temperature and coefficients of HT using two condensers and fans will be performed. This would be to develop and modify the HE in the VC HP for heating the water under different pressures, evaporator fan speeds and water inlet temperatures in the condenser, which saves energy and safeguards the environment. Work suggests that marginal modification in HP makes it cheaper for domestic applications as clean energy. 2. EXPERIMENTAL METHODOLOGY The experimental setup consists of HP mounted with two modified condensers, compressor, evaporator, capillary tube and water tank. The fluids run through the HE (condenser and evaporator) such as refrigerant (hot fluid) in the tube and water (cold fluid) in the shell side of a condenser in a counter flow mode. The refrigerant flowing inside the tubes heats up the water in the condenser. The fan mounted in the evaporator is utilized to warm up the refrigerant by taking heat from the air. The temperature and pressure instruments are fixed at the entry and exit points of HP components. Two condensers were fitted with water control valves at the inlet and outlet to control the mass stream rate of water and coupled with the HP (include an evaporator, compressor capillary tube, water tank, pump, energy meter, and fan) as showed in Figure 1. Figure 1 Open in new tabDownload slide (a) Schematic and (b) detailed layout of HP comprising two condensers and evaporator with fan. Figure 1 Open in new tabDownload slide (a) Schematic and (b) detailed layout of HP comprising two condensers and evaporator with fan. Table 1 CO2 refrigerant properties. 1 . Critical temperature °C . 31.06 . 2 Boiling point in °C −78.5 3 Critical pressure in bar 73.84 4 ODP 0 5 GWP 1 6 Molecular mass (Kg/K mol) 44.01 7 Specific heat (KJ/KgK) 0.8439 1 . Critical temperature °C . 31.06 . 2 Boiling point in °C −78.5 3 Critical pressure in bar 73.84 4 ODP 0 5 GWP 1 6 Molecular mass (Kg/K mol) 44.01 7 Specific heat (KJ/KgK) 0.8439 Open in new tab Table 1 CO2 refrigerant properties. 1 . Critical temperature °C . 31.06 . 2 Boiling point in °C −78.5 3 Critical pressure in bar 73.84 4 ODP 0 5 GWP 1 6 Molecular mass (Kg/K mol) 44.01 7 Specific heat (KJ/KgK) 0.8439 1 . Critical temperature °C . 31.06 . 2 Boiling point in °C −78.5 3 Critical pressure in bar 73.84 4 ODP 0 5 GWP 1 6 Molecular mass (Kg/K mol) 44.01 7 Specific heat (KJ/KgK) 0.8439 Open in new tab Table 2 The condensers and evaporator specifications. No. . . Condenser1 (C1) . Condenser 2 (C2) . Evaporator . 1 Configuration Counter flow. Coaxial and single pass Counter flow. Coaxial and single pass 1 HP, 1200 rpm, coaxial and single pass 2 Tube diameters and length in mm Do = 152.4, Di = 8 and L = 508 Do = 127, Di = 10 and L = 533.4 Do = 254, Di = 12, L = 330 and 3 rows 3 Centrifugal pump Energy meter 240 V, 50 Hz, single phase, AC supply 2 mm diameter contacting side to 18 mm diameter expansion Water tank 1 m3MS 4 Expansion valve 2 mm diameter contacting to 18 mm diameter expansion, water tank 1 m3MS pump Centrifugal pump No. . . Condenser1 (C1) . Condenser 2 (C2) . Evaporator . 1 Configuration Counter flow. Coaxial and single pass Counter flow. Coaxial and single pass 1 HP, 1200 rpm, coaxial and single pass 2 Tube diameters and length in mm Do = 152.4, Di = 8 and L = 508 Do = 127, Di = 10 and L = 533.4 Do = 254, Di = 12, L = 330 and 3 rows 3 Centrifugal pump Energy meter 240 V, 50 Hz, single phase, AC supply 2 mm diameter contacting side to 18 mm diameter expansion Water tank 1 m3MS 4 Expansion valve 2 mm diameter contacting to 18 mm diameter expansion, water tank 1 m3MS pump Centrifugal pump Open in new tab Table 2 The condensers and evaporator specifications. No. . . Condenser1 (C1) . Condenser 2 (C2) . Evaporator . 1 Configuration Counter flow. Coaxial and single pass Counter flow. Coaxial and single pass 1 HP, 1200 rpm, coaxial and single pass 2 Tube diameters and length in mm Do = 152.4, Di = 8 and L = 508 Do = 127, Di = 10 and L = 533.4 Do = 254, Di = 12, L = 330 and 3 rows 3 Centrifugal pump Energy meter 240 V, 50 Hz, single phase, AC supply 2 mm diameter contacting side to 18 mm diameter expansion Water tank 1 m3MS 4 Expansion valve 2 mm diameter contacting to 18 mm diameter expansion, water tank 1 m3MS pump Centrifugal pump No. . . Condenser1 (C1) . Condenser 2 (C2) . Evaporator . 1 Configuration Counter flow. Coaxial and single pass Counter flow. Coaxial and single pass 1 HP, 1200 rpm, coaxial and single pass 2 Tube diameters and length in mm Do = 152.4, Di = 8 and L = 508 Do = 127, Di = 10 and L = 533.4 Do = 254, Di = 12, L = 330 and 3 rows 3 Centrifugal pump Energy meter 240 V, 50 Hz, single phase, AC supply 2 mm diameter contacting side to 18 mm diameter expansion Water tank 1 m3MS 4 Expansion valve 2 mm diameter contacting to 18 mm diameter expansion, water tank 1 m3MS pump Centrifugal pump Open in new tab Moreover, a one-ton reciprocating compressor at 240 V, 50 Hz with temperature and compressor pressure up to 110°C and 21 bars were used. Initial reading (both pressures and temperatures) in the gauges is noted, and the inlet water temperature of a condenser is also recorded. HP was started after filling up the refrigerant to a pressure of 4 bars and allowing it to run for some time to reach a steady state. The water is supplied from the water tank to the condenser through the inlet valve using a pump. After getting the steady state, the experiment starts by recording the P&T at different system components using temperature gauges and pressure gauges. The outlet mass flow rate with a constant time interval and water temperature from the condenser is recorded. This procedure is repeated for different filling pressure (4.5 and 5 bars) of refrigerant but at a constant mass flow rate of water. The output parameters such as LMTD, coefficient of performance (COP), Q (HT) and UA (overall heat transfer coefficient) are calculated by using the recorded experimental values using available standard thermodynamic relations. The various properties of CO2 are listed in Table 1 and configurations of the experimental set up in Table 2. Table 3 Experimental results of various outputs at first condenser without preheating. . . 800 rpm . 1000 rpm . 1200 rpm . No. . mwo kg/sec . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . At 4 bars 1 0.0093 3.90 22.2 0.317 12.9 36.1 4.03 23.6 0.325 13.9 36.5 4.14 24.4 0.336 15.0 36.7 2 0.0101 3.92 21.6 0.318 13.4 35.6 4.05 22.7 0.329 14.6 35.8 4.17 23.6 0.338 15.6 36.1 3 0.011 3.96 21 0.325 14.2 35.1 4.08 22.3 0.336 15.2 35.3 4.2 22.8 0.346 16.6 35.5 4 0.012 3.99 20.4 0.329 14.9 34.6 4.1 21.5 0.341 16.0 34.8 4.23 22 0.351 17.3 35.1 At 4.5 bars 5 0.0093 3.96 24.5 0.317 12 37.6 4.1 25.2 0.332 13.0 38.1 4.21 26.6 0.341 13.9 38.7 6 0.0101 4.00 23.7 0.318 12.4 37 4.12 24.8 0.339 13.8 37.6 4.25 25.8 0.347 14.4 38.1 7 0.011 4.02 23.2 0.324 12.9 36.4 4.15 24.4 0.341 14.1 37 4.29 25.1 0.359 15.6 37.5 8 0.012 4.05 22.6 0.329 13.4 35.9 4.19 23.7 0.348 14.7 36.5 4.32 24.5 0.361 16.1 37 At 5 bars 9 0.0093 4.11 25.4 0.402 18.2 38.3 4.19 26.2 0.429 17.7 39 4.32 27.1 0.456 17.6 39.5 10 0.0101 4.12 24.9 0.415 14.8 37.9 4.22 25.6 0.447 16.4 38.5 4.35 26.5 0.467 18.5 39 11 0.011 4.13 24.6 0.426 15.9 37.4 4.24 25.1 0.46 17.4 38 4.38 25.9 0.498 20.2 38.7 12 0.012 4.14 24.2 0.447 16.5 36.9 4.29 25 0.486 19.3 37.6 4.4 25.4 0.508 20.8 38 . . 800 rpm . 1000 rpm . 1200 rpm . No. . mwo kg/sec . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . At 4 bars 1 0.0093 3.90 22.2 0.317 12.9 36.1 4.03 23.6 0.325 13.9 36.5 4.14 24.4 0.336 15.0 36.7 2 0.0101 3.92 21.6 0.318 13.4 35.6 4.05 22.7 0.329 14.6 35.8 4.17 23.6 0.338 15.6 36.1 3 0.011 3.96 21 0.325 14.2 35.1 4.08 22.3 0.336 15.2 35.3 4.2 22.8 0.346 16.6 35.5 4 0.012 3.99 20.4 0.329 14.9 34.6 4.1 21.5 0.341 16.0 34.8 4.23 22 0.351 17.3 35.1 At 4.5 bars 5 0.0093 3.96 24.5 0.317 12 37.6 4.1 25.2 0.332 13.0 38.1 4.21 26.6 0.341 13.9 38.7 6 0.0101 4.00 23.7 0.318 12.4 37 4.12 24.8 0.339 13.8 37.6 4.25 25.8 0.347 14.4 38.1 7 0.011 4.02 23.2 0.324 12.9 36.4 4.15 24.4 0.341 14.1 37 4.29 25.1 0.359 15.6 37.5 8 0.012 4.05 22.6 0.329 13.4 35.9 4.19 23.7 0.348 14.7 36.5 4.32 24.5 0.361 16.1 37 At 5 bars 9 0.0093 4.11 25.4 0.402 18.2 38.3 4.19 26.2 0.429 17.7 39 4.32 27.1 0.456 17.6 39.5 10 0.0101 4.12 24.9 0.415 14.8 37.9 4.22 25.6 0.447 16.4 38.5 4.35 26.5 0.467 18.5 39 11 0.011 4.13 24.6 0.426 15.9 37.4 4.24 25.1 0.46 17.4 38 4.38 25.9 0.498 20.2 38.7 12 0.012 4.14 24.2 0.447 16.5 36.9 4.29 25 0.486 19.3 37.6 4.4 25.4 0.508 20.8 38 Open in new tab Table 3 Experimental results of various outputs at first condenser without preheating. . . 800 rpm . 1000 rpm . 1200 rpm . No. . mwo kg/sec . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . At 4 bars 1 0.0093 3.90 22.2 0.317 12.9 36.1 4.03 23.6 0.325 13.9 36.5 4.14 24.4 0.336 15.0 36.7 2 0.0101 3.92 21.6 0.318 13.4 35.6 4.05 22.7 0.329 14.6 35.8 4.17 23.6 0.338 15.6 36.1 3 0.011 3.96 21 0.325 14.2 35.1 4.08 22.3 0.336 15.2 35.3 4.2 22.8 0.346 16.6 35.5 4 0.012 3.99 20.4 0.329 14.9 34.6 4.1 21.5 0.341 16.0 34.8 4.23 22 0.351 17.3 35.1 At 4.5 bars 5 0.0093 3.96 24.5 0.317 12 37.6 4.1 25.2 0.332 13.0 38.1 4.21 26.6 0.341 13.9 38.7 6 0.0101 4.00 23.7 0.318 12.4 37 4.12 24.8 0.339 13.8 37.6 4.25 25.8 0.347 14.4 38.1 7 0.011 4.02 23.2 0.324 12.9 36.4 4.15 24.4 0.341 14.1 37 4.29 25.1 0.359 15.6 37.5 8 0.012 4.05 22.6 0.329 13.4 35.9 4.19 23.7 0.348 14.7 36.5 4.32 24.5 0.361 16.1 37 At 5 bars 9 0.0093 4.11 25.4 0.402 18.2 38.3 4.19 26.2 0.429 17.7 39 4.32 27.1 0.456 17.6 39.5 10 0.0101 4.12 24.9 0.415 14.8 37.9 4.22 25.6 0.447 16.4 38.5 4.35 26.5 0.467 18.5 39 11 0.011 4.13 24.6 0.426 15.9 37.4 4.24 25.1 0.46 17.4 38 4.38 25.9 0.498 20.2 38.7 12 0.012 4.14 24.2 0.447 16.5 36.9 4.29 25 0.486 19.3 37.6 4.4 25.4 0.508 20.8 38 . . 800 rpm . 1000 rpm . 1200 rpm . No. . mwo kg/sec . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . At 4 bars 1 0.0093 3.90 22.2 0.317 12.9 36.1 4.03 23.6 0.325 13.9 36.5 4.14 24.4 0.336 15.0 36.7 2 0.0101 3.92 21.6 0.318 13.4 35.6 4.05 22.7 0.329 14.6 35.8 4.17 23.6 0.338 15.6 36.1 3 0.011 3.96 21 0.325 14.2 35.1 4.08 22.3 0.336 15.2 35.3 4.2 22.8 0.346 16.6 35.5 4 0.012 3.99 20.4 0.329 14.9 34.6 4.1 21.5 0.341 16.0 34.8 4.23 22 0.351 17.3 35.1 At 4.5 bars 5 0.0093 3.96 24.5 0.317 12 37.6 4.1 25.2 0.332 13.0 38.1 4.21 26.6 0.341 13.9 38.7 6 0.0101 4.00 23.7 0.318 12.4 37 4.12 24.8 0.339 13.8 37.6 4.25 25.8 0.347 14.4 38.1 7 0.011 4.02 23.2 0.324 12.9 36.4 4.15 24.4 0.341 14.1 37 4.29 25.1 0.359 15.6 37.5 8 0.012 4.05 22.6 0.329 13.4 35.9 4.19 23.7 0.348 14.7 36.5 4.32 24.5 0.361 16.1 37 At 5 bars 9 0.0093 4.11 25.4 0.402 18.2 38.3 4.19 26.2 0.429 17.7 39 4.32 27.1 0.456 17.6 39.5 10 0.0101 4.12 24.9 0.415 14.8 37.9 4.22 25.6 0.447 16.4 38.5 4.35 26.5 0.467 18.5 39 11 0.011 4.13 24.6 0.426 15.9 37.4 4.24 25.1 0.46 17.4 38 4.38 25.9 0.498 20.2 38.7 12 0.012 4.14 24.2 0.447 16.5 36.9 4.29 25 0.486 19.3 37.6 4.4 25.4 0.508 20.8 38 Open in new tab Table 4 Experimental results of various outputs at second condenser without preheating. . . 800 rpm . 1000 rpm . 1200 rpm . No. . mwo kg/sec . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . At 4 bars 1 0.0093 3.98 22.6 0.369 15.2 36.1 4.12 23.8 0.39 16.7 36.5 4.23 24.4 0.415 18.4 36.8 2 0.0101 4.01 21.8 0.377 16.0 35.6 4.14 22.8 0.404 17.8 36 4.27 23.7 0.426 19.6 36.1 3 0.011 4.04 21.2 0.388 16.7 35 4.16 22.3 0.413 18.6 35.4 4.3 23.2 0.436 20.8 35.8 4 0.012 4.06 20.5 0.398 17.6 34.5 4.21 21.7 0.427 19.7 35 4.33 22.5 0.451 22.1 35.3 At 4.5 bars 5 0.0093 4.05 24.4 0.42 15.6 38.7 4.18 25.5 0.445 17.3 39.4 4.3 26.6 0.468 19.2 40 6 0.0101 4.09 23.5 0.425 16.6 38 4.21 24.8 0.461 18.4 38.9 4.33 25.5 0.481 20.4 39.4 7 0.011 4.1 22.9 0.436 17.4 37.4 4.24 24.1 0.489 20.3 38.6 4.36 24.9 0.505 22.0 39 8 0.012 4.14 22.3 0.461 18.9 37.1 4.27 23.6 0.492 20.9 37.7 4.39 24.3 0.52 23.2 38.4 At 5 bars 9 0.0093 4.17 27.1 0.462 16.7 39.9 4.3 27.6 0.491 17.7 40.6 4.42 27.8 0.502 18.3 40.9 10 0.0101 4.2 26 0.491 18.1 39.7 4.33 26.5 0.511 19.2 40.1 4.46 27.3 0.511 19.9 40.3 11 0.011 4.24 25.1 0.515 19.6 39.3 4.36 25.8 0.533 20.7 39.5 4.48 26.5 0.533 21.2 39.7 12 0.012 4.26 24 0.515 20.0 38.4 4.4 25.4 0.536 21.0 38.6 4.53 26 0.547 23.1 39 . . 800 rpm . 1000 rpm . 1200 rpm . No. . mwo kg/sec . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . At 4 bars 1 0.0093 3.98 22.6 0.369 15.2 36.1 4.12 23.8 0.39 16.7 36.5 4.23 24.4 0.415 18.4 36.8 2 0.0101 4.01 21.8 0.377 16.0 35.6 4.14 22.8 0.404 17.8 36 4.27 23.7 0.426 19.6 36.1 3 0.011 4.04 21.2 0.388 16.7 35 4.16 22.3 0.413 18.6 35.4 4.3 23.2 0.436 20.8 35.8 4 0.012 4.06 20.5 0.398 17.6 34.5 4.21 21.7 0.427 19.7 35 4.33 22.5 0.451 22.1 35.3 At 4.5 bars 5 0.0093 4.05 24.4 0.42 15.6 38.7 4.18 25.5 0.445 17.3 39.4 4.3 26.6 0.468 19.2 40 6 0.0101 4.09 23.5 0.425 16.6 38 4.21 24.8 0.461 18.4 38.9 4.33 25.5 0.481 20.4 39.4 7 0.011 4.1 22.9 0.436 17.4 37.4 4.24 24.1 0.489 20.3 38.6 4.36 24.9 0.505 22.0 39 8 0.012 4.14 22.3 0.461 18.9 37.1 4.27 23.6 0.492 20.9 37.7 4.39 24.3 0.52 23.2 38.4 At 5 bars 9 0.0093 4.17 27.1 0.462 16.7 39.9 4.3 27.6 0.491 17.7 40.6 4.42 27.8 0.502 18.3 40.9 10 0.0101 4.2 26 0.491 18.1 39.7 4.33 26.5 0.511 19.2 40.1 4.46 27.3 0.511 19.9 40.3 11 0.011 4.24 25.1 0.515 19.6 39.3 4.36 25.8 0.533 20.7 39.5 4.48 26.5 0.533 21.2 39.7 12 0.012 4.26 24 0.515 20.0 38.4 4.4 25.4 0.536 21.0 38.6 4.53 26 0.547 23.1 39 Open in new tab Table 4 Experimental results of various outputs at second condenser without preheating. . . 800 rpm . 1000 rpm . 1200 rpm . No. . mwo kg/sec . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . At 4 bars 1 0.0093 3.98 22.6 0.369 15.2 36.1 4.12 23.8 0.39 16.7 36.5 4.23 24.4 0.415 18.4 36.8 2 0.0101 4.01 21.8 0.377 16.0 35.6 4.14 22.8 0.404 17.8 36 4.27 23.7 0.426 19.6 36.1 3 0.011 4.04 21.2 0.388 16.7 35 4.16 22.3 0.413 18.6 35.4 4.3 23.2 0.436 20.8 35.8 4 0.012 4.06 20.5 0.398 17.6 34.5 4.21 21.7 0.427 19.7 35 4.33 22.5 0.451 22.1 35.3 At 4.5 bars 5 0.0093 4.05 24.4 0.42 15.6 38.7 4.18 25.5 0.445 17.3 39.4 4.3 26.6 0.468 19.2 40 6 0.0101 4.09 23.5 0.425 16.6 38 4.21 24.8 0.461 18.4 38.9 4.33 25.5 0.481 20.4 39.4 7 0.011 4.1 22.9 0.436 17.4 37.4 4.24 24.1 0.489 20.3 38.6 4.36 24.9 0.505 22.0 39 8 0.012 4.14 22.3 0.461 18.9 37.1 4.27 23.6 0.492 20.9 37.7 4.39 24.3 0.52 23.2 38.4 At 5 bars 9 0.0093 4.17 27.1 0.462 16.7 39.9 4.3 27.6 0.491 17.7 40.6 4.42 27.8 0.502 18.3 40.9 10 0.0101 4.2 26 0.491 18.1 39.7 4.33 26.5 0.511 19.2 40.1 4.46 27.3 0.511 19.9 40.3 11 0.011 4.24 25.1 0.515 19.6 39.3 4.36 25.8 0.533 20.7 39.5 4.48 26.5 0.533 21.2 39.7 12 0.012 4.26 24 0.515 20.0 38.4 4.4 25.4 0.536 21.0 38.6 4.53 26 0.547 23.1 39 . . 800 rpm . 1000 rpm . 1200 rpm . No. . mwo kg/sec . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . COP . LMTD in °C . Q in KW . UA*10−3 in KW/k . Two in °C . At 4 bars 1 0.0093 3.98 22.6 0.369 15.2 36.1 4.12 23.8 0.39 16.7 36.5 4.23 24.4 0.415 18.4 36.8 2 0.0101 4.01 21.8 0.377 16.0 35.6 4.14 22.8 0.404 17.8 36 4.27 23.7 0.426 19.6 36.1 3 0.011 4.04 21.2 0.388 16.7 35 4.16 22.3 0.413 18.6 35.4 4.3 23.2 0.436 20.8 35.8 4 0.012 4.06 20.5 0.398 17.6 34.5 4.21 21.7 0.427 19.7 35 4.33 22.5 0.451 22.1 35.3 At 4.5 bars 5 0.0093 4.05 24.4 0.42 15.6 38.7 4.18 25.5 0.445 17.3 39.4 4.3 26.6 0.468 19.2 40 6 0.0101 4.09 23.5 0.425 16.6 38 4.21 24.8 0.461 18.4 38.9 4.33 25.5 0.481 20.4 39.4 7 0.011 4.1 22.9 0.436 17.4 37.4 4.24 24.1 0.489 20.3 38.6 4.36 24.9 0.505 22.0 39 8 0.012 4.14 22.3 0.461 18.9 37.1 4.27 23.6 0.492 20.9 37.7 4.39 24.3 0.52 23.2 38.4 At 5 bars 9 0.0093 4.17 27.1 0.462 16.7 39.9 4.3 27.6 0.491 17.7 40.6 4.42 27.8 0.502 18.3 40.9 10 0.0101 4.2 26 0.491 18.1 39.7 4.33 26.5 0.511 19.2 40.1 4.46 27.3 0.511 19.9 40.3 11 0.011 4.24 25.1 0.515 19.6 39.3 4.36 25.8 0.533 20.7 39.5 4.48 26.5 0.533 21.2 39.7 12 0.012 4.26 24 0.515 20.0 38.4 4.4 25.4 0.536 21.0 38.6 4.53 26 0.547 23.1 39 Open in new tab Table 5 Experimental results of various outputs at first and second condensers with preheated water at 36°C. No. . Fan speed in rpm . mwo in Kg/sec . COP at 4.5 bars . COP at 5 bars . Q at 4.5 bars . Q at 5 bars . LMTD at 4.5 bars . LMTD at 5 bars . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . 1 800 0.0093 4.19 4.23 4.31 4.34 0.56 0.62 0.64 0.70 16.82 17.81 16.93 17.67 2 0.0101 4.22 4.27 4.35 4.38 0.58 0.62 0.66 0.71 15.30 15.90 15.53 16.30 3 0.011 4.25 4.30 4.37 4.39 0.59 0.63 0.67 0.76 13.53 14.44 14.61 15.56 4 0.012 4.27 4.32 4.38 4.40 0.61 0.65 0.69 0.77 12.79 13.37 14.33 15.00 5 1000 0.0093 4.30 4.34 4.41 4.45 0.59 0.65 0.69 0.73 18.73 19.73 18.89 20.05 6 0.0101 4.32 4.37 4.43 4.46 0.60 0.66 0.70 0.74 16.96 17.89 17.49 18.50 7 0.011 4.36 4.40 4.46 4.49 0.61 0.66 0.71 0.78 15.19 16.50 16.15 17.53 8 0.012 4.38 4.43 4.49 4.52 0.63 0.67 0.77 0.82 14.26 14.98 15.65 16.51 9 1200 0.0093 4.41 4.44 4.52 4.55 0.63 0.67 0.73 0.76 20.07 21.27 21.05 21.73 10 0.0101 4.44 4.48 4.55 4.57 0.64 0.68 0.75 0.78 18.68 20.33 19.44 20.30 11 0.011 4.46 4.49 4.57 4.61 0.65 0.69 0.75 0.81 17.74 18.49 18.58 18.98 12 0.012 4.49 4.53 4.60 4.64 0.66 0.70 0.80 0.84 15.92 16.84 16.71 17.49 No. . Fan speed in rpm . mwo in Kg/sec . COP at 4.5 bars . COP at 5 bars . Q at 4.5 bars . Q at 5 bars . LMTD at 4.5 bars . LMTD at 5 bars . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . 1 800 0.0093 4.19 4.23 4.31 4.34 0.56 0.62 0.64 0.70 16.82 17.81 16.93 17.67 2 0.0101 4.22 4.27 4.35 4.38 0.58 0.62 0.66 0.71 15.30 15.90 15.53 16.30 3 0.011 4.25 4.30 4.37 4.39 0.59 0.63 0.67 0.76 13.53 14.44 14.61 15.56 4 0.012 4.27 4.32 4.38 4.40 0.61 0.65 0.69 0.77 12.79 13.37 14.33 15.00 5 1000 0.0093 4.30 4.34 4.41 4.45 0.59 0.65 0.69 0.73 18.73 19.73 18.89 20.05 6 0.0101 4.32 4.37 4.43 4.46 0.60 0.66 0.70 0.74 16.96 17.89 17.49 18.50 7 0.011 4.36 4.40 4.46 4.49 0.61 0.66 0.71 0.78 15.19 16.50 16.15 17.53 8 0.012 4.38 4.43 4.49 4.52 0.63 0.67 0.77 0.82 14.26 14.98 15.65 16.51 9 1200 0.0093 4.41 4.44 4.52 4.55 0.63 0.67 0.73 0.76 20.07 21.27 21.05 21.73 10 0.0101 4.44 4.48 4.55 4.57 0.64 0.68 0.75 0.78 18.68 20.33 19.44 20.30 11 0.011 4.46 4.49 4.57 4.61 0.65 0.69 0.75 0.81 17.74 18.49 18.58 18.98 12 0.012 4.49 4.53 4.60 4.64 0.66 0.70 0.80 0.84 15.92 16.84 16.71 17.49 Open in new tab Table 5 Experimental results of various outputs at first and second condensers with preheated water at 36°C. No. . Fan speed in rpm . mwo in Kg/sec . COP at 4.5 bars . COP at 5 bars . Q at 4.5 bars . Q at 5 bars . LMTD at 4.5 bars . LMTD at 5 bars . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . 1 800 0.0093 4.19 4.23 4.31 4.34 0.56 0.62 0.64 0.70 16.82 17.81 16.93 17.67 2 0.0101 4.22 4.27 4.35 4.38 0.58 0.62 0.66 0.71 15.30 15.90 15.53 16.30 3 0.011 4.25 4.30 4.37 4.39 0.59 0.63 0.67 0.76 13.53 14.44 14.61 15.56 4 0.012 4.27 4.32 4.38 4.40 0.61 0.65 0.69 0.77 12.79 13.37 14.33 15.00 5 1000 0.0093 4.30 4.34 4.41 4.45 0.59 0.65 0.69 0.73 18.73 19.73 18.89 20.05 6 0.0101 4.32 4.37 4.43 4.46 0.60 0.66 0.70 0.74 16.96 17.89 17.49 18.50 7 0.011 4.36 4.40 4.46 4.49 0.61 0.66 0.71 0.78 15.19 16.50 16.15 17.53 8 0.012 4.38 4.43 4.49 4.52 0.63 0.67 0.77 0.82 14.26 14.98 15.65 16.51 9 1200 0.0093 4.41 4.44 4.52 4.55 0.63 0.67 0.73 0.76 20.07 21.27 21.05 21.73 10 0.0101 4.44 4.48 4.55 4.57 0.64 0.68 0.75 0.78 18.68 20.33 19.44 20.30 11 0.011 4.46 4.49 4.57 4.61 0.65 0.69 0.75 0.81 17.74 18.49 18.58 18.98 12 0.012 4.49 4.53 4.60 4.64 0.66 0.70 0.80 0.84 15.92 16.84 16.71 17.49 No. . Fan speed in rpm . mwo in Kg/sec . COP at 4.5 bars . COP at 5 bars . Q at 4.5 bars . Q at 5 bars . LMTD at 4.5 bars . LMTD at 5 bars . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . 1 800 0.0093 4.19 4.23 4.31 4.34 0.56 0.62 0.64 0.70 16.82 17.81 16.93 17.67 2 0.0101 4.22 4.27 4.35 4.38 0.58 0.62 0.66 0.71 15.30 15.90 15.53 16.30 3 0.011 4.25 4.30 4.37 4.39 0.59 0.63 0.67 0.76 13.53 14.44 14.61 15.56 4 0.012 4.27 4.32 4.38 4.40 0.61 0.65 0.69 0.77 12.79 13.37 14.33 15.00 5 1000 0.0093 4.30 4.34 4.41 4.45 0.59 0.65 0.69 0.73 18.73 19.73 18.89 20.05 6 0.0101 4.32 4.37 4.43 4.46 0.60 0.66 0.70 0.74 16.96 17.89 17.49 18.50 7 0.011 4.36 4.40 4.46 4.49 0.61 0.66 0.71 0.78 15.19 16.50 16.15 17.53 8 0.012 4.38 4.43 4.49 4.52 0.63 0.67 0.77 0.82 14.26 14.98 15.65 16.51 9 1200 0.0093 4.41 4.44 4.52 4.55 0.63 0.67 0.73 0.76 20.07 21.27 21.05 21.73 10 0.0101 4.44 4.48 4.55 4.57 0.64 0.68 0.75 0.78 18.68 20.33 19.44 20.30 11 0.011 4.46 4.49 4.57 4.61 0.65 0.69 0.75 0.81 17.74 18.49 18.58 18.98 12 0.012 4.49 4.53 4.60 4.64 0.66 0.70 0.80 0.84 15.92 16.84 16.71 17.49 Open in new tab Table 6 Experimental results of various outputs at first and second condensers with preheated water at 36°C. No. . Fan speed in rpm . mwo in Kg/sec . UA in KW/K at 4.5 bar . UA in KW/K at 5 bars . Two in °C at 4.5 bar . Two in °C at 4.5 bar . . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . 1 800 0.0093 0.028 0.028 0.030 0.032 46.56 47.89 48.36 49.92 2 0.0101 0.030 0.030 0.034 0.035 45.74 46.77 47.59 48.85 3 0.011 0.033 0.034 0.036 0.038 44.88 45.83 46.45 48.00 4 0.012 0.037 0.038 0.041 0.044 44.15 44.90 45.82 47.43 5 1000 0.0093 0.032 0.033 0.036 0.036 47.37 48.63 49.57 50.69 6 0.0101 0.034 0.037 0.040 0.040 46.18 47.46 48.48 49.52 7 0.011 0.040 0.040 0.045 0.044 45.22 46.34 47.48 48.95 8 0.012 0.044 0.045 0.049 0.050 44.36 45.16 47.13 48.18 9 1200 0.0093 0.038 0.037 0.043 0.043 48.17 49.11 50.67 51.51 10 0.0101 0.042 0.043 0.048 0.048 47.02 48.09 49.61 50.29 11 0.011 0.048 0.048 0.051 0.052 45.97 47.06 48.27 49.51 12 0.012 0.052 0.052 0.056 0.056 45.06 45.94 47.84 48.82 No. . Fan speed in rpm . mwo in Kg/sec . UA in KW/K at 4.5 bar . UA in KW/K at 5 bars . Two in °C at 4.5 bar . Two in °C at 4.5 bar . . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . 1 800 0.0093 0.028 0.028 0.030 0.032 46.56 47.89 48.36 49.92 2 0.0101 0.030 0.030 0.034 0.035 45.74 46.77 47.59 48.85 3 0.011 0.033 0.034 0.036 0.038 44.88 45.83 46.45 48.00 4 0.012 0.037 0.038 0.041 0.044 44.15 44.90 45.82 47.43 5 1000 0.0093 0.032 0.033 0.036 0.036 47.37 48.63 49.57 50.69 6 0.0101 0.034 0.037 0.040 0.040 46.18 47.46 48.48 49.52 7 0.011 0.040 0.040 0.045 0.044 45.22 46.34 47.48 48.95 8 0.012 0.044 0.045 0.049 0.050 44.36 45.16 47.13 48.18 9 1200 0.0093 0.038 0.037 0.043 0.043 48.17 49.11 50.67 51.51 10 0.0101 0.042 0.043 0.048 0.048 47.02 48.09 49.61 50.29 11 0.011 0.048 0.048 0.051 0.052 45.97 47.06 48.27 49.51 12 0.012 0.052 0.052 0.056 0.056 45.06 45.94 47.84 48.82 Open in new tab Table 6 Experimental results of various outputs at first and second condensers with preheated water at 36°C. No. . Fan speed in rpm . mwo in Kg/sec . UA in KW/K at 4.5 bar . UA in KW/K at 5 bars . Two in °C at 4.5 bar . Two in °C at 4.5 bar . . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . 1 800 0.0093 0.028 0.028 0.030 0.032 46.56 47.89 48.36 49.92 2 0.0101 0.030 0.030 0.034 0.035 45.74 46.77 47.59 48.85 3 0.011 0.033 0.034 0.036 0.038 44.88 45.83 46.45 48.00 4 0.012 0.037 0.038 0.041 0.044 44.15 44.90 45.82 47.43 5 1000 0.0093 0.032 0.033 0.036 0.036 47.37 48.63 49.57 50.69 6 0.0101 0.034 0.037 0.040 0.040 46.18 47.46 48.48 49.52 7 0.011 0.040 0.040 0.045 0.044 45.22 46.34 47.48 48.95 8 0.012 0.044 0.045 0.049 0.050 44.36 45.16 47.13 48.18 9 1200 0.0093 0.038 0.037 0.043 0.043 48.17 49.11 50.67 51.51 10 0.0101 0.042 0.043 0.048 0.048 47.02 48.09 49.61 50.29 11 0.011 0.048 0.048 0.051 0.052 45.97 47.06 48.27 49.51 12 0.012 0.052 0.052 0.056 0.056 45.06 45.94 47.84 48.82 No. . Fan speed in rpm . mwo in Kg/sec . UA in KW/K at 4.5 bar . UA in KW/K at 5 bars . Two in °C at 4.5 bar . Two in °C at 4.5 bar . . C1 . C2 . C1 . C2 . C1 . C2 . C1 . C2 . 1 800 0.0093 0.028 0.028 0.030 0.032 46.56 47.89 48.36 49.92 2 0.0101 0.030 0.030 0.034 0.035 45.74 46.77 47.59 48.85 3 0.011 0.033 0.034 0.036 0.038 44.88 45.83 46.45 48.00 4 0.012 0.037 0.038 0.041 0.044 44.15 44.90 45.82 47.43 5 1000 0.0093 0.032 0.033 0.036 0.036 47.37 48.63 49.57 50.69 6 0.0101 0.034 0.037 0.040 0.040 46.18 47.46 48.48 49.52 7 0.011 0.040 0.040 0.045 0.044 45.22 46.34 47.48 48.95 8 0.012 0.044 0.045 0.049 0.050 44.36 45.16 47.13 48.18 9 1200 0.0093 0.038 0.037 0.043 0.043 48.17 49.11 50.67 51.51 10 0.0101 0.042 0.043 0.048 0.048 47.02 48.09 49.61 50.29 11 0.011 0.048 0.048 0.051 0.052 45.97 47.06 48.27 49.51 12 0.012 0.052 0.052 0.056 0.056 45.06 45.94 47.84 48.82 Open in new tab Figure 2 Open in new tabDownload slide COP vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 2 Open in new tabDownload slide COP vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 3 Open in new tabDownload slide LMTD vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 3 Open in new tabDownload slide LMTD vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. The same procedure is repeated for preheated water at 36°C supplied to the condenser at different filling pressure (4.5 and 5 bars) of refrigerant and constant mass flow rate of water (mwo). 3. RESULTS AND DISCUSSION HP’s various important results with two condensers having different mwo rates with varying fan speeds and pressures are measured. Multiple graphs of performance characteristics w.r.t. above said parameters involved in HP’s design and performance utilizing CO2 refrigerant are plotted. Tables 3–6 of appendix A show the experimental results of various outputs for two condensers with and without preheating. 3.1. Inlet water without preheating at 28°C 3.1.1. Results of COP Figure 2a–c shows how COP varies w.r.t. mass flow rate for condensers 1 and 2 at various pressures and the different fan speeds of the evaporator (800, 1000 and 1200 rpm), respectively. The percentage increase of COP is observed with increased pressure and mass flow rate corresponding to the evaporator’s fan speed rise. The maximum and minimum COPs were recorded as 4.4 and 3.9 in condenser 1 and 4.53 and 3.98 in condenser 2. Higher COP is observed at a speed of 1200 rpm compared to the other two speeds as effective heat extraction in the evaporator increases HP’s performance in the cyclic process. The COP is seen in the range of 3.9 to 4.53 for the condensers, while condenser 2 showed the highest COP (4.53) for 1200 rpm fan speed at 5 bar pressure and condenser 1 exhibited a minimum value of COP (3.9) at 4 bars and 800 rpm speed. The increase in the inlet temperature of water supplied to the condenser also increased COP. By comparing the COP of the condensers, condenser 2 showed enhanced performance as more HT area is exposed in the HE. The increasing value of pressure raised the COP of the HP because of the increase in thermodynamic properties (temperature, volume) of the refrigerant to transfer more heat to cooling fluid in the condenser. Condenser 2 has tube diameter is higher than that of a condenser 1 where the refrigerant flow increases with large HT area and it rises pressure. Pressure rise changes temperature difference and hence COP. 3.1.2. Results of LMTD The following Figure 3a–c shows variation of LMTD against mass flow rate for condenser 1 and condenser 2 at varying pressures and fan speed of evaporator (800, 1000 and 1200 rpm), respectively. The increasing fan speed increased the value of LMTD because fan speed enables quick HT between air and refrigerant. Whereas, LMTD was decreasing with the increased mass flow of water for a given evaporator fan speed as water moving with high speed cannot absorb the heat and temperature difference decreases. Maximum LMTD is observed at fan speed of 1200 rpm with maximum and minimum LMTD of 27.1 and 20.4°C in condenser 1 and 27.8 and 20.5°C in condenser 2, respectively. The increasing value of pressure increased LMTD for various fan speeds. The HE with a counter flow method is suitable for improved results. The end temperature difference at both sides of the condenser is found more which increased the value of LMTD. The LMTD fall in the range of 20.4 to 27.8°C in the condensers with maximum of LMTD (27.8°C) for condenser 2 at 5 bars pressure and 800 rpm fan speed, whereas minimum LMTD (20.4°C) for condenser 1 at 4 bars pressure and 1200 rpm fan speed were found. The pressure rise increases the LMTD of the HP because of increase in thermodynamic properties (temperature, volume) of the refrigerant increases the end temperature difference between the two sides of the counter flow HE resulting in an increase of COP. 3.1.3. Results of HT The Figure 4a–c shows the results of Q against mass flow rates for two condensers at various pressures and three fan speeds of evaporator, respectively. The maximum Q of 0.508 kW found in condenser 1 and 0.547 kW in condenser 2, respectively, were observed at maximum fan speed of 1200 rpm. The HT, Q, is obtained in the range of 0.508 to 0.547 kW for two condensers showing condenser 2, maximum of Q (0.547 kW) is achieved for 1200 rpm fan speed at a pressure of 5 bars and minimum of Q (0.317 kW) was seen at 4 bars pressure and 800 rpm fan speed. Maximum amount of Q was observed in the second condenser for four reasons. Firstly, its tube diameter is 25% higher than the tube of a first condenser. Secondly, the refrigerant flow increased in the second condenser with large area of HT. Thirdly, the fan speed enabled quick HT between air and refrigerant. Finally, HT increased with rise in pressure as it changes temperature difference. Figure 4 Open in new tabDownload slide Q of HP vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 4 Open in new tabDownload slide Q of HP vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 5 Open in new tabDownload slide U of HP vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm (b) 1000 rpm and (c) 1200 rpm. Figure 5 Open in new tabDownload slide U of HP vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm (b) 1000 rpm and (c) 1200 rpm. Figure 6 Open in new tabDownload slide Two of HP vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 6 Open in new tabDownload slide Two of HP vs mwo (at 28°C) at pressures of 4, 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 7 Open in new tabDownload slide COP of HP vs mwo (at 36°C) at pressures of 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c)1200 rpm. Figure 7 Open in new tabDownload slide COP of HP vs mwo (at 36°C) at pressures of 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c)1200 rpm. Figure 8 Open in new tabDownload slide LMTD of HP vs mwo (at 36°C) at pressures of 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 8 Open in new tabDownload slide LMTD of HP vs mwo (at 36°C) at pressures of 4.5 and 5 bars for (a) 800 rpm, (b) 1000 rpm and (c) 1200 rpm. Figure 9 Open in new tabDownload slide Q of HP vs mwo (at 36°C) at pressures of 4.5 and 5 bars for (a) 800, (b) 1000 rpm and (c) 1200 rpm. Figure 9 Open in new tabDownload slide Q of HP vs mwo (at 36°C) at pressures of 4.5 and 5 bars for (a) 800, (b) 1000 rpm and (c) 1200 rpm. Figure 10 Open in new tabDownload slide UA of HP vs mwo (at 36°C) at pressures of 4.5 & 5 bars for (a) 800, (b) 1000 rpm and (c)1200 rpm. Figure 10 Open in new tabDownload slide UA of HP vs mwo (at 36°C) at pressures of 4.5 & 5 bars for (a) 800, (b) 1000 rpm and (c)1200 rpm. Figure 11 Open in new tabDownload slide Two of HP vs mwo (at 36°C) at pressures of 4.5 and 5 bars for (a) 800, (b) 1000 and (c) 1200 rpm. Figure 11 Open in new tabDownload slide Two of HP vs mwo (at 36°C) at pressures of 4.5 and 5 bars for (a) 800, (b) 1000 and (c) 1200 rpm. 3.1.4. Results of overall HT coefficient, UA The Figure 5a–c represents the results of overall HT coefficient versus mass flow rate for condensers 1 and 2 for different pressures and fan speeds (800, 1000 and 1200 rpm), respectively. The UA found increasing with increased water mass flow rate and fan speed. The maximum and minimum UAs are observed in the range of 0.0208 and 0.012 for condenser 1 and 0.0152 and 0.0231 for condenser 2, respectively. Improved UA is noticed at the fan speed of 1200 rpm in the range of 0.0208 to 0.0231 KW/K for the condensers. In condenser 2, maximum UA (0.0231) is achieved for 1200 rpm fan speed of the evaporator at a pressure of 5 bars. In condenser 1, minimum overall HT coefficient (0.012) is recorded at 4 bar pressure and 1000 rpm fan speed. With the increasing heat transfer rate, the water specific heat (Cp) remains invariable. Therefore, overall heat transfer coefficient (U) increases. Concerning to the law of conservation of energy, increase in water mass flow rate (mwo) increases the overall HT coefficient. 3.1.5. Results of outlet water temperature The Figure 6a–c shows results of outlet water temperature versus mass flow rate for different fan speed of evaporator (800, 1000 and1200 rpm) and pressures (4, 4.5 and 5 bars). The condenser fed the water from the tank through water pump. The water outlet temperature increases with pressure rise but decreases with rise in water mass flow rate. The maximum and minimum outlet water temperatures were found 39.5 and 34.6°C (condenser 1) and 41 and 34.5°C (condenser 2), respectively. The water outlet temperature was in the range of 35 to 40°C for the condensers. Maximum of 41°C is achieved at 1200 rpm fan speed and a pressure of 5 bars in condenser 2 and minimum of 34.5°C at 800 rpm and a pressure of 4 bars in the condenser 1. The UA is found maximum in condenser 2 because of the higher diameter of tube than condenser 1, HT rate increases. Also, by the time it reaches condenser 2, heat absorbed by water is from the two stages of condensation where most of heat is absorbed. Heat absorbed by water is proportional to U alone as Cp is constant for a given mwo. The fan speed of evaporator also increases the HT and outlet water temperature due to quick HTs between air and refrigerant. 3.2. Preheating of water at 36°C 3.2.1. Results of COP The Figure 7a–c shows variation of COP against mass flow rate for different fan speed of evaporator and pressures with water preheated to 36°C, respectively. The behavior of COP remains same without preheating w.r.t. pressure, mass flow rate and fan speed. The maximum and minimum COPs measured were 4.60 and 4.19 (condenser 1) and 4.64 and 4.23 (condenser 2), respectively, with a range of 4.19 to 4.64 for both the condensers. Condenser 2 gives maximum COP of 4.64 for 1200 rpm fan speed at a pressure of 5 bars and condenser 1 gives minimum of 4.19 at 4.5 bars and 800 rpm. By preheating the inlet water at 36°C, COP increased in both condensers but condenser 2 showed better COP. COP of the HP increases in thermodynamic properties (temperature, volume) of the refrigerant to transfer more heat to cooling fluid in the condenser. 3.2.2. LMTD variation (preheated water at 36°C) The Figure 8a–c shows LMTD of HP versus mass flow rate at different fan speed of evaporator and pressures with water preheated to 36°C for two condensers. The variation of LMTD remains same as seen in LMTD without preheating w.r.t fan speed and pressures. The maximum and minimum LMTDs were 21.27 and 12.79°C (condenser 1) and 21.73 and 13.37°C (condenser 2), respectively, with a range of 12.79 to 21.73°C for both condensers. LMTD decreases in case of inletting preheated water at 36°C and it improves the performance of the HP. 3.2.3. Results of HT, Q, for preheated water at 36°C The Figure 9a–c presents the results of HT against mass flow rate for both condenser at different fan speed and pressures. The pattern of Q remains same as in the case of without preheating w.r.t. fan speeds and pressures. The HT is recorded in the range of 0.564 to 0.843 KW for both condensers with maximum Q of 0.843for condenser 2 at high speed and high pressure and minimum Q of 0.564 is recorded at low pressure and fan speed. Maximum Q is found in condenser 2 because of its higher tube size than condenser 1 for refrigerant flow, which increases area of HT. The Q is enhanced from 0.584 to 0.843 KW by preheating the inlet water at 36°C under similar conditions. 3.2.4. Results of overall HT coefficient, UA for preheated water at 36°C The Figure 10a–c shows the overall HT coefficient against mass flow rate for two condensers at different fan speed and pressures. After preheating inlet water at 36°C, the UA is improved from 0.028KW/k to 0.052 KW/K. The UA is recorded around 0.028 to 0.056KW/K for both the condensers. The condenser 2 showed maximum UA of 0.056 at 1200 rpm and 5 bars, whereas in condenser 1, minimum UA of 0.028 is found at 4.5 bars and 800 rpm. 3.2.5. Results of outlet water temperature for preheated water at 36°C The Figure 11a–c depicts the results of water outlet temperature against the mass flow rate for both the condenser at different pressures and fan speeds, respectively. The range of water outlet temperature is recorded 44.1°C to 50.67°C for the condensers. In condenser 2, maximum Two of 50.67 is found out at 1200 rpm and a pressure of 5 bars and in condenser 1, the minimum of 44.15 is found at 4 bars and 800 rpm. The Two is improved from 49.3°C to 51.5°C by preheating the inlet water. Referring to Tables 3 and 4 and comparing with the results of conventional refrigerant (R32)-based HP, the heat transferring capacity increases in transcritical CO2 despite R32, which has rich refrigerant mixture. A refrigerant at a high pressure shows a low specific volume, which increase mwo and the heat transferring capacity. COP changed from 4.4 to 4.53 in the transcritical CO2 HP test compared to R32-based HP. CO2 improved COP of HP almost twice the HP working on R32 (COP = 2.3 to 2.85) under the similar conditions. Whereas, LMTD for condenser varied in the range of 20–26°C for R32-based HP, and those for CO2-based HP is 22.2 to 26°C [31]. COP decreases as composition of R32 increases but in case of transcritical CO2-based HP, COP improves. pollution also decreased in transcritical CO2-based HP as there are no CFC or HCFC, etc. This shows the advantages of using transcritical CO2-based HP. 4. CONCLUSION CO2 has been used as a natural refrigerant, non-toxic, non-combustible, safe working refrigerant, low-cost, non-destructive, steady and appropriate for a wide variety of working conditions. The CFC compounds have been first substitutes for higher ODP. The GWP and ODP are 1 and 0, respectively, for CO2 refrigerant, and it is best for HP compared with other refrigerants. In the case of water heating, HP is an effective air-based heating system designed and fabricated, having two condensers and with as well as without preheating of inlet water. This HP would replace the electrical heating system. The performance based on the experimental investigation of the CO2 HP to heat the water by different fan speeds of the evaporator, pressure and mass flow rate leads to the following conclusions: With increasing, mass flow rate, the COP, UA and Q increased, but LMTD decreased for increasing fan speed and pressure. The outlet water temperature increases as the mass flow rate decreases. Improved COP, HT and outlet water temperature are observed from the experiment by preheating water at 36°C in other places. Whereas in the Kingdom of Saudi Arabia, more than 9 months in a year, the temperature is above 36°C does not even need preheating of water. Maximum Q, Two COP and the overall HT coefficient were found by adding the second condenser. The CO2 refrigerant performance shows better efficiency and reduced environmental impact when compared with globally used R32 refrigerant. Therefore, CO2 refrigerant is the best replacement for globally used artificial refrigerants. If this technology is improved and implemented on a large scale, the TCC-CO2-based HP device can compete with other non-renewable energy-based devices (namely, Solar, cow dung gas plant, wind power). Authors contribution All the authors have equally contributed to the work. Acknowledgement Not applicable. References [1] Olson DA . 2000 . Heat transfer of supercritical carbon dioxide flowing in a cooled horizontal tube . Report No. 6496 . Gaithersburg : NIST Publications . Google Scholar Google Preview OpenURL Placeholder Text WorldCat COPAC [2] Zhang Z , Tian L, Tong L et al. Choked flow characteristics of subcritical refrigerant flowing through converging-diverging nozzles . Entropy 2014 ; 16 : 5810 – 21 . doi: 10.3390/e16115810 . Google Scholar Crossref Search ADS WorldCat [3] Sarkar J . Transcritical CO2 heat pump simulation model and validation for simultaneous cooling and heating . Int J Mech Mechatronics Eng 2008 ; 2 : 249 – 54 . Google Scholar OpenURL Placeholder Text WorldCat [4] Elbel S , Hrnjak P. Flash gas bypass for improving the performance of transcritical R744 systems that use microchannel evaporators . Int J Refrig 2004 ; 27 : 724 – 35 . doi: 10.1016/j.ijrefrig.2004.07.019 . Google Scholar Crossref Search ADS WorldCat [5] Robinson DM , Groll EA. Efficiencies of transcritical CO2 cycles with and without an expansion turbine . Int J Refrig 1998 ; 21 : 577 – 89 . Google Scholar Crossref Search ADS WorldCat [6] Kim MH , Pettersen J, Bullard CW. Fundamental process and system design issues in CO2 vapor compression systems . Prog Energy Combust Sci 2004 ; 30 : 119 – 74 . doi: 10.1016/j.pecs.2003.09.002 . Google Scholar Crossref Search ADS WorldCat [7] Dnyanesh J , Malhar J. Prototype development and testing of air source heat pump water heater for Indian climate . Int J Mech Prod Eng 2015 ; 3 : 81 – 5 . Google Scholar OpenURL Placeholder Text WorldCat [8] Lorentzen G . Revival of carbon dioxide as a refrigerant . Int J Refrig 1994 ; 17 : 292 – 301 . doi: 10.1016/0140-7007(94)90059-0 . Google Scholar Crossref Search ADS WorldCat [9] Lorentzen G , Pettersen J. A new, efficient and environmentally benign system for car air-conditioning . Int J Refrig 1993 ; 16 : 4 – 12 . doi: 10.1016/0140-7007(93)90014-Y. Google Scholar Crossref Search ADS WorldCat [10] Jahar S . Review on cycle modifications of transcritical CO2 refrigeration and heat pump systems . J Adv Res Mech Eng 2015 ; 1 : 22 – 9 . Google Scholar OpenURL Placeholder Text WorldCat [11] Lorentzen G . The use of natural refrigerants: a complete solution to the CFC/HCFC predicament . Int J Refrig 1995 ; 18 : 190 – 7 . doi: 10.1016/0140-7007(94)00001-E . Google Scholar Crossref Search ADS WorldCat [12] Aprea C , Maiorino A. An experimental evaluation of the transcritical CO 2 refrigerator performances using an internal heat exchanger .doi: 10.1016/j.ijrefrig.2007.12.016 . [13] Wang J , Hihara E. Study on carbon dioxide gas cooler heat transfer process under supercritical pressures . Int J Energy Res 2002 ; 26 : 1237 – 51 . doi: 10.1002/er.847. Google Scholar Crossref Search ADS WorldCat [14] Pitla SS , Groll EA, Ramadhyani S. Convective heat transfer from in-tube cooling of turbulent supercritical carbon dioxide: part 2—experimental data and numerical predictions . HVAC R Res 2001 ; 7 : 367 – 82 .doi: 10.1080/10789669.2001.10391281. Google Scholar Crossref Search ADS WorldCat [15] Pettersen J , Hafner A, Skaugen G et al. Development of compact heat exchangers for CO2 air-conditioning systems . Int J Refrig 1998 ; 21 : 180 – 93 . doi: 10.1016/S0140-7007(98)00013-9 . Google Scholar Crossref Search ADS WorldCat [16] Süß J , Kruse H. Efficiency of the indicated process of CO2-compressors . Int J Refrig 1998 ; 21 : 194 – 201 . doi: 10.1016/s0140-7007(98)00022-x . Google Scholar OpenURL Placeholder Text WorldCat Crossref [17] Gnielinski V . New equations for heat and mass transfer in the turbulent flow in pipes and channels . STIA 1975 ; 41 : 8 – 16 . Google Scholar OpenURL Placeholder Text WorldCat [18] Kim S , Kim M. Experiment and simulation on the performance of an autocascade refrigeration system using carbon dioxide as arefrigerant . Int J Refrig 2002 ; 25 : 1093 – 101 . doi: 10.1016/S0140-7007(01)00110-4 [19] Mozurkewich G , Greenfield ML, Schneider WF et al. Simulated performance and cofluid dependence of a CO2-cofluid refrigeration cycle with wet compression . Int J Refrig 2002 ; 25 : 1123 – 36 . doi: 10.1016/S0140-7007(02)00004-X . Google Scholar Crossref Search ADS WorldCat [20] Sivasakthivel T , Siva Kumar Reddy KK. Ozone layer depletion and its effects: a review . Int J Environ Sci Dev 2011 ; 2 : 30 – 7 . Google Scholar OpenURL Placeholder Text WorldCat [21] Mason M . Protecting the ozone layer: the United Nations history . Eur Environ 2003 ; 13 : 183 – 4 . doi: 10.1002/eet.311 . Google Scholar Crossref Search ADS WorldCat [22] Summary for Policymakers—Global Warming of 1.5°C . https://www.ipcc.ch/sr15/chapter/spm/ (24 August 2020, date last accessed). [23] Basavaraju MG , Ravi P, Sreenivas Rao KV. Importance of using CO2 refrigerant in refrigeration, air conditioning and heat pump . Int J Mech Eng Res 2015 ; 5 : 135 – 42 . Google Scholar OpenURL Placeholder Text WorldCat [24] Babarinde TO , Ohunakin OS, Adelekan DS et al. Experimental study of LPG and R134a refrigerants in vapor compression refrigeration . Int J Energy Clean Environ 2015 ; 16 : 71 – 80 . doi: 10.1615/InterJEnerCleanEnv.2016015644 . Google Scholar Crossref Search ADS WorldCat [25] Sukhikh AA , Antanenkova IS, Kuznetsov VN et al. Experimental and numerical investigations of heat regeneration process efficiency in a heat pump with a mixture of refrigerants . Int J Energy Clean Environ 2016 ; 17 : 209 – 22 . doi: 10.1615/InterJEnerCleanEnv.2017019186 . Google Scholar Crossref Search ADS WorldCat [26] Lv Y , Bi J, Yan J. State-of-the-art in low carbon community . Int J Energy Clean Environ 2018 ; 19 : 175 – 200 . doi: 10.1615/InterJEnerCleanEnv.2018025415 . Google Scholar Crossref Search ADS WorldCat [27] Kamble D , Gawali BS. Analysis of a rectangular microchannel using the R-22 refrigerant in forced convection heat transfer condition . Int J Energy Clean Environ 2019 ; 20 : 1 – 21 . doi: 10.1615/InterJEnerCleanEnv.2016015640 . Google Scholar Crossref Search ADS WorldCat [28] Zhou J , Wang J, Yan Z et al. Development and application of a microchannel heat exchanger for the heat pump . Int J Energy Clean Environ 2018 ; 19 : 137 – 41 . doi: 10.1615/INTERJENERCLEANENV.2018020983 . Google Scholar Crossref Search ADS WorldCat [29] Mkanda FX , Mutimba S, Mwanyoka I et al. Wood-saving stoves: an opportunity for households to switch technology and contribute to reduction of deforestation and carbon dioxide emission. The case of Kilimanjaro Region, Tanzania . Int J Energy Clean Environ 2014 ; 15 : 35 – 46 .doi: 10.1615/InterJEnerCleanEnv.2015014117 . Google Scholar Crossref Search ADS WorldCat [30] Karnaukh VV , Birukov AB, Mazur VA et al. Comparative analysis of different refrigerants used in a high-temperature vapor-compression heat pump . Int J Energy Clean Environ 2017 ; 18 : 161 – 74 . doi: 10.1615/InterJEnerCleanEnv.2017022821 . Google Scholar Crossref Search ADS WorldCat [31] Kim M , Kim MS, Kim Y. Experimental study on the performance of a heat pump system with refrigerant mixtures’ composition change . Energy 2004 ; 29 : 1053 – 68 . Google Scholar Crossref Search ADS WorldCat © The Author(s) 2021. Published by Oxford University Press. This is an Open Access article distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/4.0/), which permits unrestricted reuse, distribution, and reproduction in any medium, provided the original work is properly cited.
TI - Performance characteristics of a heat pump using renewable transcritical CO2 refrigerant for heating water
JF - International Journal of Low-Carbon Technologies
DO - 10.1093/ijlct/ctab009
DA - 2021-02-23
UR - https://www.deepdyve.com/lp/oxford-university-press/performance-characteristics-of-a-heat-pump-using-renewable-Zo5ZkzNGnX
SP - 1
EP - 1
VL - Advance Article
IS - 
DP - DeepDyve
ER -