OVERVIEW
Review of related literature
This study has been experimentally supported and theoretically based on following national and international related literature reviews:
Buytaert et al [1] investigated the effectiveness of microchannel evaporative CO2 cooling for enhancing the LHCb vertex detector Their research indicates that this cooling method can satisfy the stringent thermal efficiency requirements for the upgraded LHCb VELO detector, achieving a low operational temperature of -27°C, a high power density of 12.9 W, and a minimal mass Additionally, the study demonstrates that temperature gradients of less than 4°C can be attained within the module.
We are addressing the issue of high pressure resistance and reliability to operate in the detector vacuum.
Li et al [2] conducted a study on CO2 gas removal for micro passive direct methanol fuel cells (DMFC), introducing a novel configuration featuring super hydrophobic lateral venting microchannels This design facilitates the direct release of CO2 gas from the MEA diffusion layer, significantly reducing the gas's passage through the anode flow field and preventing accumulation in the methanol solution reservoir The enclosed DMFC with this lateral venting configuration demonstrated enhanced performance at high current densities and low methanol concentrations, attributed to reduced anode methanol-transport resistance and decreased anode concentration polarization The I-V curve from constant-current-density discharging tests indicated a more stable output voltage, supported by observations of fewer bubbles on the anode This innovative lateral venting approach presents a promising avenue for optimizing DMFCs and warrants further investigation into different channel arrangements and applications in larger flow fields.
Roychowdhury et al conducted research on the conjugate heat transfer in the steam reforming of ethanol within microchannel systems, concluding that complete ethanol conversion is achievable at flue gas temperatures above approximately 1400 °C However, the thermal efficiency stabilizes beyond a certain flue gas inlet temperature, highlighting the need to identify optimal heating conditions for improved energy utilization Despite achieving 100% ethanol conversion, hydrogen yield is influenced by methanation and water-gas shift reactions, which vary with system temperature At lower reforming fluid temperatures (~600 K), high methane concentrations lead to low hydrogen yields, while increasing the mixture temperature enhances carbon monoxide and hydrogen concentrations Conversely, at very high temperatures, hydrogen yield declines due to the reverse water-gas shift reaction, indicating that an optimal mixture temperature of around 1000 K maximizes hydrogen yield from the reformer system.
Gaudillere et al [4] investigated CO2 hydrogenation using Ru/Ce-based catalysts on highly ordered microchannelled 3YSZ monoliths created through freeze-casting Their research evaluated the performance of these catalysts for CO2 methanation at temperatures of 450°C and 500°C in two configurations The findings revealed that Ru/Ce-based catalysts effectively facilitate the reaction, with the monolith configuration reducing CO production and enhancing CH4 selectivity Increasing the H2 content in the inlet stream was found to shift the RWGS reaction equilibrium, further improving CH4 yield Notably, the study indicated no significant pressure drop from carbon deposition, supporting the viability of freeze-cast porous samples as catalyst supports These results underscore the importance of precise catalyst composition control in optimizing reaction yields The research suggests potential enhancements for improving CH4 selectivity, minimizing carbon deposition, and increasing CO2 conversion, including modifications to pore microstructure, better catalyst precursor infiltration control, and optimization of metal loading.
Kim et al conducted a study on the impact of micro-grooves on the two-phase pressure drop of CO2 in a minichannel tube, revealing significant findings regarding the behavior of CO2 flow in such configurations.
+ Using the hydraulic diameter, the pressure drop of CO2 in the grooved minichannel could be successfully estimated with conventional correlations for smooth channels.
The Blasius correlation accurately predicted the pressure drop in both channels during the liquid phase Notably, the grooved tube exhibited a pressure drop approximately 1.3 times greater than that of the smooth tube, attributed to its smaller hydraulic diameter.
+ In the two-phase experiments, the pressure drop of the grooved channel was found to be 1.1-1.45 times greater than that of the smooth channel.
Experiments conducted under constant heat flux conditions demonstrated that pressure drop data obtained in adiabatic conditions can accurately predict the actual pressure drop in the evaporator, even as the quality changes.
Most two-phase pressure drop models exhibit a mean absolute error ranging from 17% to 35% for smooth channels and 13% to 32% for grooved channels Notably, the model developed by Mishima and Hibiki in 1995 demonstrated the most accurate results.
Liang et al [6] conducted an investigation into the flow boiling heat transfer characteristics of CO2 in horizontal mini-tubes Their experimental study analyzed the effects of heat flux, mass flux, and saturation temperature on the heat transfer coefficient and dryout characteristics By comparing the experimental results with theoretical prediction models, they drew significant conclusions regarding the behavior of CO2 in these conditions.
Compared to traditional refrigerants, CO2 exhibits superior thermal properties, leading to a higher heat transfer coefficient in minichannels, particularly in the low to medium vapor quality range However, as vapor quality increases, dryout occurs, causing a significant decline in heat exchange performance and a reduction in the average heat transfer coefficient throughout the heat transfer process.
The heat flux plays a crucial role in influencing the heat transfer coefficient during working conditions An increase in heat flux significantly enhances nucleate boiling heat transfer, leading to a rise in the heat transfer coefficient.
The increase in heat flux during working conditions accelerates the dryout process and affects vapor quality at both the beginning and end stages As mass flux rises, the initial vapor quality of dryout decreases Additionally, higher saturation temperatures facilitate the occurrence of dryout, resulting in lower starting vapor quality.
Cheng's model demonstrates superior prediction accuracy when compared to existing theories, effectively forecasting the heat transfer coefficient prior to the onset of dryout The model achieves a prediction deviation of within ±30% and an overall accuracy rate of 77.1%.
Jiang et al [7] explored the enhancement of CO2 dissolution and sweep efficiency in saline aquifers through micro bubble (MB) CO2 injection Their study involved supercritical CO2 injection into a Berea sandstone sample at a rate of 0.05 mL/min under reservoir conditions, analyzed using X-ray CT The experiments revealed that MB CO2, compared to normal bubble (NB) CO2, showed accelerated dissolution in the early stages of injection, with a notable 18.7% increase in dissolution efficiency at 0.02 PV Additionally, MB CO2 improved pore space utilization due to its higher permeability in low porosity regions, achieving a 4.9% enhancement in efficiency These findings indicate that MB CO2 sequestration is an effective method for CO2 storage.
CO2 in low-porosity regions by enhancing dissolution and improving pore space utilization.
Liang et al [8] investigated the enhancement of micro direct methanol fuel cell (DMFC) performance through in situ CO2 removal using a novel anode flow field featuring superhydrophobic degassing channels Their findings revealed that the power densities of the N-Serpentine and N-Spiral micro DMFCs improved by over 30% and 90%, respectively, compared to the C-Serpentine and C-Spiral configurations This significant performance boost is primarily due to the more effective removal of CO2.
Research background
Meeting the energy demands for heating and cooling is crucial for a nation's economic and scientific growth While heating applications are predominantly utilized in European countries like England, Russia, and Sweden, cooling systems have gained global popularity In recent years, industries have focused on enhancing cooling efficiency, leading to innovations such as Variable Refrigerant Volume (VRV) systems and eco-friendly air conditioning solutions These advancements have been successfully implemented in various practical applications.
The indoor unit, also known as the evaporator, plays a crucial role in refrigeration systems by facilitating heat transfer from the environment to the refrigerant To enhance the cooling capacity of these evaporators, numerous experimental studies and simulations have been conducted, aiming to improve the overall performance of the refrigeration system.
CO 2 convective heat transfer coefficient in microchannel evaporator was carried out.
Authors’ achievement
To enhance heat transfer efficiency, it is essential to investigate the convective heat transfer coefficient of CO2 in microchannel heat exchangers Improving heat transfer efficiency involves various strategies, including minimizing heat loss to the environment, maintaining and replacing damaged components, and regularly checking device functions A thorough understanding of heat transfer fundamentals and the conceptual principles of thermodynamics is crucial for advancing the study of convective heat transfer coefficient improvements.
Methodology
Review methodology: other related studies are compiled in order to set up the research’s direction and objective.
Experimental methodology: obtained measurement data will undergo theoretical calculation, experimental analysis and correlation.
Numerical simulation methodology: simulation model will have been prepared viaCAD drawing and the whole process will be executed with Comsol Multiphysics software.
Research objective
The firstly prioritized objective in this study is expected to be heat transfer coefficient in CO2 microchannel evaporator The research area is limited to HCMUTE laboratory.
INTRODUCTION
Microchannel and micro heat exchanger
The term microchannel came from microtechnology developed in electrical and electronic engineering All micro channels have the diameters of below 1000 μm and are mostly applied in microfluidics and heat transfer [28].
2.1.1 A brief history of micro technology
In 1981, Tuckerman et al demonstrated both theoretically and experimentally that reducing the dimensions of a conventional plate-fin liquid-cooled heat sink to approximately 50 μm in channel width, while operating within the laminar flow regime and integrating it into silicon chips, resulted in a remarkable 20-fold increase in specific thermal conductance and over 1000-fold enhancement in volumetric heat removal during laboratory tests.
In Tuckerman's study, it was noted that water entered the device at room temperature, was rapidly heated to 135°C within approximately 2.5 seconds, maintained at that temperature for around 2.5 seconds, and then cooled back down in about 2.5 seconds, exiting at a temperature only slightly above its initial state, demonstrating the high effectiveness of the heat exchanger.
Investigation of micro-scale thermal devices by Tuckerman et al is motivated by the single phase internal flow correlation for convective heat transfer: h = (2.1) where:
+ d: Hydraulic diameter of the channel or duct (m).
Internal laminar flows in tubes or channels yield a constant Nusselt number, with experimental values indicating Nuc = 3.657 for constant wall temperature and Nuc = 4.364 for constant heat flux in round tubes In flat parallel plates, the Nusselt number is 8.23 Laminar flow is characterized by a Reynolds number proportional to the hydraulic diameter, and as the hydraulic diameter decreases, the heat transfer coefficient increases Consequently, increasing the hydraulic diameter to 10 or 100 μm in forced convection can significantly enhance the heat transfer coefficient.
Micro HX, or micro-scale heat exchangers, are specialized devices designed for fluid flows in microchannels with diameters under 1 mm Typically constructed from metal or ceramic, these heat exchangers are utilized in high-performance applications such as aircraft gas turbine engines, heat pumps, and air conditioning systems They share a classification system with traditional macro heat exchangers and can be categorized into three types based on the flow patterns of the working fluids: parallel, counter, and perpendicular flow.
2.1.3 Micro HX’s advantages and disadvantages compared to those of conventional HX a Advantages
+ Substantially higher performance than that of other heat exchangers due to having many micro-scale channels.
+ Lower weight reduces transportation and support structure cost.
+ Lower cost for fabrication materials. b Disadvantages
The main disadvantage of small-scale hydraulic diameter in micro heat exchangers is the resulting pressure loss, which disrupts the uniform flow of refrigerant within the channels To enhance the performance of micro heat exchangers, it is crucial to balance the convective heat transfer coefficient with the pressure loss equation.
+ Microchannels are sometimes fairly long and absorb most of the heat along the first section of the channel, which makes them less able to absorb heat along later sections.
+ Wall roughness is considered to be some of factor in calculating heat transfer coefficient.
Thermodynamic properties of R744
Carbon dioxide (CO2) is a colorless, odorless, and non-toxic gas that is heavier than air, posing a suffocation risk in enclosed spaces When dissolved in water, it forms carbonic acid, and at low temperatures, it can solidify into dry ice Typically existing as a gas, CO2 can change to liquid or solid states under high pressure or low temperatures Importantly, carbon dioxide is considered environmentally friendly as it does not contribute to global warming or ozone depletion.
Solid and liquid carbon dioxide (R744) are vital refrigerants, particularly in the food industry, due to their favorable thermal properties for various applications such as refrigeration, cooling, and heating After considering the climate impact of R134a, liquid CO2 has been repurposed as a more environmentally friendly alternative Compressed at pressures up to 130 bar, R744 systems necessitate robust components designed for mass production Notably, 90% of operation logs for vehicle air conditioning systems utilizing R744 demonstrate superior performance compared to R134a, especially at temperatures of 50°C or higher Given its eco-friendly nature, R744 presents a promising opportunity to replace HFCs in vehicles, supermarkets, heat pump economizers, and other heating and cooling applications in the near future.
Heat transfer and heat transfer coefficient fundamentals
The term heat transfer stands in the field of physic and thermal engineering as a discipline of following mechanisms: thermal conduction, thermal convection and thermal radiation.
Heat conduction is the process where kinetic energy is exchanged between particles across the boundary of two systems This transfer of heat occurs from an object at a higher temperature to one at a lower temperature until thermal equilibrium is achieved.
Heat convection occurs when a fluid moves, carrying its matter with it, and can be categorized into two types: natural convection, where fluid movement is driven by natural forces without external influence, and forced convection, where mechanical means are employed to induce fluid flow.
Thermal radiation is the transfer of energy through a vacuum or transparent mediums such as solids, fluids, or gases, and it occurs via photons in electromagnetic waves, adhering to established physical laws.
The heat transfer coefficient, also known as the film coefficient or film effectiveness, is a crucial parameter in thermodynamics and mechanics, representing the relationship between heat flux and the temperature difference driving heat flow A detailed calculation methodology for this coefficient will be explored in the subsequent chapter.
Air conditioning fundamentals
Air conditioning is the process of removing heat and moisture from indoor spaces to create a comfortable environment for occupants While primarily aimed at enhancing comfort for humans and animals, air conditioning also plays a crucial role in cooling and dehumidifying areas with heat-generating electronic devices, such as computer servers and power amplifiers Additionally, it is essential for preserving delicate items like artwork by maintaining optimal temperature and humidity levels.
Air conditioners primarily operate on the single-stage refrigeration cycle, also known as the vapor compression cycle This traditional cooling method relies on the forced circulation and phase change of a refrigerant between gas and liquid states to effectively transfer heat.
Sensible heat refers to the heat transferred by a thermodynamic system that results in a change in temperature, affecting certain macroscopic variables, while keeping others, like volume and pressure, constant.
Latent heat is an energy form contributing to the state change of a materials or an object without changing its temperature As the term goes, latent heat always stays
Latent heat plays a crucial role in refrigeration systems by supplying or extracting the necessary heat to transform solids into liquids or vapors, and liquids into vapors By balancing mass and energy equations, the theoretical principles of latent heat are validated, highlighting its importance in effective temperature regulation.
Numerical simulation with COMSOL Multiphysics
The COMSOL Multiphysics software offers several different formulations for solving turbulent flow problems: the L-VEL, algebraic yPlus, Spalart-Allmaras, k-ε, k-ω, low Reynolds number k-ε, SST, and v2-f turbulence models.
The L-VEL and algebraic yPlus turbulence models compute the eddy viscosity using algebraic expressions based only on the local fluid velocity and the distance to the closest wall.
The Spalart-Allmaras model introduces an extra variable for undamped kinematic eddy viscosity, making it a low Reynolds number model This model effectively captures the complete flow field, extending down to the solid wall.
The k-ε model solves for two variables: k, the turbulence kinetic energy; and ε (epsilon), the rate of dissipation of turbulence kinetic energy.
The k-ω model is similar to the k-ε model, but it solves for ω (omega) — the specific rate of dissipation of kinetic energy.
The low Reynolds number k-ε model extends the traditional k-ε model by eliminating the need for wall functions, enabling it to solve for flow throughout the entire domain While it retains many advantages of the original k-ε model, it typically necessitates a finer mesh due to its low Reynolds number characteristics, which affect turbulence not just at the walls but across the entire flow field.
The SST model integrates the k-ε model for free stream conditions with the k-ω model for wall proximity, making it a low Reynolds number model This approach has established the SST model as a preferred choice in computational fluid dynamics.
DESIGN AND METHODOLOGY
Experimental system
The measurement was conducted on an R744 air conditioning system, which is monitored through an electric control box and features a micro-scale evaporator This closed experimental system consists of vapor and liquid pipelines made of copper, along with four key components: the compressor, gas cooler, throttling valve, and evaporator The measurement process was enhanced by various instruments, which will be detailed in the subsequent section.
Figure 3.1 Schematic diagram of the experimental system
The experimental system operates similarly to the vapor compression cycle, as illustrated in Figure 3.1, with pressure gauges located at specific points (1, 1’, 2, 3, and 4) The circulating CO2 exits the evaporator as a saturated vapor and absorbs heat from the evaporator's fan, transforming into a superheated vapor This vapor then enters the compressor, where it is isentropically compressed to a higher pressure superheated vapor The refrigerant subsequently undergoes isobaric cooling in the gas cooler, followed by passing through a throttling valve, leading to isenthalpic expansion and resulting in flash evaporation as it exits the valve as a low-pressure vapor Finally, the refrigerant re-enters the evaporator, absorbs heat from the air, and completes its cycle by exiting as a saturated vapor.
3.1.1 Semi-hermetic single-stage reciprocating compressor
The experimental compressor is manufactured by Dorin Innovation Company, Italy. The device’s technical data and other related information are listed in Table 3.1, Figure 3.2 and 3.3.
Table 3.1 Specifications of Dorin compressor
Max operating current 4,4A (380 - 420V/3ph/50Hz (Y)) Locked rotor current 20A(380 - 420V/3ph/50Hz (Y))
Table 3.2 Legends and symbols in Figure 3.3
DL (discharge line) Discharge service valve
SL (suction line) Suction service valve
LPSV Low pressure safety valve
HPSV High pressure safety valve
The experimental micro evaporator is manufactured and fabricated out of aluminum with specifications shown in Figure 3.4 and 3.5.
The experimental gas cooler’s cover is fabricated out of stainless steel with its aluminum heat exchanger shown in Figure 3.6.
Figure 3.6 The front (left) and back (right) of the experimental gas cooler
During the experiment, a throtling valve (Figure 3.7) will be adjusted every 15 minutes to obtain required valve’s inlet pressure value (P3 in this case).
The electric control box used in the experiment monitors the compressor, evaporator, and gas cooler fans, as illustrated in Figure 3.8 Additionally, it regulates the evaporator fan speed during measurements to ensure accurate experimental data collection.
Measurement instruments
To obtain measurement data at the lowest errors, indicators, sensors and sorts of meter are highly required.
Pressure values are recorded via these instruments with installation location shown in Figure 3.9.
Figure 3.9 p1, p2 (left) and P3, p4 (right) pressure gauges
Final calculation and analysis should utilize most of temperature and humidity readings from the indicators shown in the following figures (3.10 and 3.11). a Thermo-hygrometer
Thermo hygrometers are useful for measurements of humidity In this study, only room temperature and relative humidity values are to be collected via this instrument.
Figure 3.10 (From left) Thermo hygrometer, DTTS, digital temperature meter and temperature humidity meter b Digital thermometer thermocouple sensor (DTTS)
The Digital Temperature Measurement System (DTTS) operates on principles similar to conventional temperature meters, but it offers enhanced accuracy with reduced error margins This instrument effectively measures the temperature values at both the inlet and outlet of the throttling valve, ensuring precise readings for optimal performance.
All locations in the Figure 3.1 of the experimental system with temperature of 50 o C or below are installed with digital temperature meter. d Temperature humidity meter
In this study, evaporator inlet and outlet relative humidity are measured by this instrument.
In this research, an anemometer and a clamp meter are used to collect evaporator’s air velocity as shown in Figure 3.11.
Figure 3.11 Anemometer (left) and clamp meter (right)
Experimental data collection process
+ System data will be collected when not inoperative through all measurement instruments discussed in the previous section.
If the pressure gauge readings at the designated points in Figure 3.1 fall below 52 bar, it is essential to perform a refrigerant charge to ensure that all readings achieve the target value of 52 bar.
Figure 3.12 Powering the control box (left) and other components (right)
To initiate system operation, switch on the power supply of the control box and press the START button This action powers the fans of the evaporator and gas cooler After a delay of 10 minutes, the process is repeated to power the compressor.
Figure 3.13 Switching on evaporator fan (left), gas cooler fan (middle) and compressor
+ Evaporator fan’s speed is adjusted to 4.5 m/s after calculating the average result via placing the anemometer in front of the fan cage (Figure 3.14).
Figure 3.14 Adjusting fan speed (left) and collecting air velocity values (right)
+ The same process is implemented for the temperature humidity meter but in addition of placing temperature humidity meter’s bulb at the surface of the micro heat exchanger.
The throttling valve is gradually adjusted to modify the valve inlet pressure (P3), beginning at 74 bar and increasing in increments of 1 bar Throughout this process, all instrument readings are meticulously monitored and documented every 15 minutes for each value of P3.
+ Compressor power supply is firstly switched off; the process is repeated for evaporator’s and gas cooler’s fan after 10 minutes; power supply for control box is switched off eventually.
+ Experimental data obtained after system operation will be carefully analyzed and technically tested via theoretical calculation and numerical simulation in the following chapter.
Numerical simulation process
Simulation model (geometrical object) in this study is a single pass taken from the micro-scale evaporator with actual size shown in Figure 3.4.
Comsol Multiphysics software was initially launched, followed by the setup of the model environment in a specific sequence: Model Wizard → 3D → Heat Transfer → Conjugate Heat Transfer → Turbulent Flow, k-ε To ensure optimal simulation conditions, Study 1 was configured as Stationary (see Figure 3.15).
The object drawing had been prepared with Autodesk Inventor software (Figure 3.16).Due to poor simulation support facility, the pass length was limited to 10 cm instead of
The object was then imported from Autodesk Inventor to field Geometry 1 in Comsol Multiphysics (Figure 3.17).
Figure 3 17 Geometrical object imported from Autodesk Inventor software
Boundary conditions were defined with input parameters in Table 3.3 After that, the numerical simulation was immediately started after selecting Compute.
Table 3.3 Input parameters for boundary conditions
Multiphysics Disable non-isothermal Flow 1
User-controlled mesh with size
Size 1 (free tetrahedral 1) extremely coarse extremely coarse extremely coarse extremely coarse (custom: minimum size element size 0.45)
RESULTS AND DISCUSSION
Measurement data
P3 74.0 75.0 76.0 77.0 78.0 bar p4 50.2 50.1 50.1 49.9 49.9 bar tambient 31.5 31.5 31.6 31.2 31.8 o C tair outlet 20.7 20.7 20.5 20.2 19.8 o C φair outlet 96.5 95.4 92.9 93.8 93.3 % teva inlet 13.1 12.7 12.6 12.0 11.3 o C teva outlet 19.7 19.5 19.3 18.8 18.3 o C tvalve inlet 30.9 31.1 31.2 31.4 31.6 o C tvalve outlet 12.80 12.40 12.10 11.80 11.00 o C troom 29.0 28.8 28.5 28.0 27.7 o C φair inlet 67.4 66.3 65.0 64.4 63.8 % tsuction 26.9 26.7 26.2 26.1 26.0 o C tdischarge 68.0 69.0 69.9 71.7 72.8 o C
The relationship between temperature and pressure is depicted in Figure 4.1, highlighting that a decrease in suction pressure leads to an increase in discharge pressure This phenomenon can be attributed to the additional work required by the compressor, as outlined in Table 4.5, resulting from a reduced refrigerant supply at the compressor inlet Notably, the gradual reduction in evaporator inlet and room temperatures indicates an enhancement in system performance, which will be further explored in the subsequent sections of this paper.
T em pe ra tu re ( oC ) P re ss u re (b ar )
Figure 4.1 Room and evaporator inlet temperature in refer to suction and discharge pressure
Figure 4.2 shows experimental system’s operational cycle in refer to obtained tabular values listed in Table 4.1 at throttling valve inlet pressure (P3) of 74 bar.
Figure 4.2 System cycle graphically illustrated by p-h diagram
Table 4.2 shows enthalpy values at designated points theoretically interpolated and extrapolated via Table A (appendix).
Table 4.1 Enthalpy values at designated points
Value Unit h1 417.5340 417.5340 417.5340 417.7730 417.7730 kJ/kg h1’ 445.2770 444.8928 443.9322 444.1828 443.9927 kJ/kg h2 468.1345 468.1391 467.7094 468.1185 468.1916 kJ/kg h4 = h3 346.3544 342.9392 338.9322 336.6467 334.5752 kJ/kg
From this section onward, further calculation and analysis on CTHC will be experimentally discussed.
CHTC’s affection on system performance investigation methodology
Thermodynamics as well as other energy and mass balance related equations are conceptually applied to help investigate the structural affection of microchannels on system performance in terms of CHTC.
The air conditioning system utilized in this experiment was previously installed and referenced in related studies, making the consistent data from those studies valuable for the current investigation.
+ Evaporator’s air velocity (measured by anemometer): vair = 4.5 m/s.
+ Surface area of evaporator fan: Ffan = 0.05027 m 2
+ Surface area of evaporator: Feva = 2.5 m 2
+ Thickness measured for an aluminum layer: δ = 0.4 mm = 0.0004 m.
+ Heat transfer coefficient of air blow in evaporator: αair = 106 w/m 2 o K.
+Specific heat of air: = 1.021 kJ/kg o K.
+ Mass flow rate of vapor CO2 flowing through compressor: = 70 kg/h = 0.0194 kg/s [29].
+ Latent heat of vaporization of water: hfg = 2501 kJ/kg.
To accurately assess the overall cooling load, it is essential to evaluate the overall heat transfer coefficient This analysis aims to achieve the most precise results by mathematically examining cooling load quantities related to air flow and CO2 levels Consequently, this chapter will conclude with a discussion of two sets of results, focusing on air and CO2 contributions to cooling load.
Overall cooling load in CO2 terms is calculated from 2 following equations:
Overall heat transfer conductivity in CO2 terms is analyzed after mathematically rearranging (4.2):
Since the air blow direction is perpendicular to that of CO2 flow, can be determined via LMTD methodology as below [26]:
+ , : inlet and outlet air blow temperature, o C;
+ , : inlet and outlet CO2 temperature in the evaporator, o C;
Figure 4.3 Air to CO2 heat transfer pattern is then substituted back to (4.3) to obtain CHTC is eventually determined by rearranging general equation of overall heat transfer coefficient (4.7, 4.8):
= , w/m 2 o K (4.8) b investigation in terms of air blow
When analyzing air blow, it is essential to consider relative humidity and the associated physical laws The total cooling load is determined by the combination of latent heat (Ql) and sensible heat (Qs), necessitating an exploration of the scientific relationships between humidity and air properties.
Overall cooling load and its constituents are determined as below:
Ql = Gair × hfg × (W1 - W2), kW (4.11) where:
+ Gkk: mass flow rate of air, kg/s The average air density (ρair, kg/m 3 ) between evaporator inlet and outlet ar temperature is determined by interpolation;
Gair = vair × Ffan × ρair, kg/s (4.12)
+ W1, W2: evaporator inlet and outlet absolute humidity obtained via Daikin Psychrometric Program Viewer, kg/kgd air.
Similar calculation and investigation are chronologically applied in order to generally determine CHTC:
Q = kair × Feva × = 2.5 × kair × , kW (4.13) kair = , W/m 2 o K (4.14)
According to air conditioning and refrigeration concept, system’s coefficient of performance (COP) can be determined as below:
Further investigation on some more quantities can be discussed for more accurate results and analyses. a Heat flux
Since the CHTC investigation has been implemented in terms of air blow and CO2, overall heat flux must be determined respectively:
Practical compressor’s power consumption is obtained via below equation:
+ I: electric current passing through compressor, A.
+ cos (φ) = 0.9: power factor [26]. c Gas cooler heat load
Similar calculation as done with the evaporator cooling load is also applied to obtain gas cooler heat load:
Experimental results analysis
Table 4.3, 4.4 and 4.5 show the achievement of above calculation and theoretical investigation on CHTC and related quantities.
Table 4.2 Calculation database in CO2 terms
Correlations between those quantities were graphically analyzed as well as carried out in terms of numerical simulation to reach conclusion on CHTC behaviors and system’s further development.
Table 4.3 Calculation database in air blow terms
A significant increase in Convective Heat Transfer Coefficient (CHTC) was noted alongside a slight decrease in suction pressure, as illustrated in Figure 4.3 The peak CHTC values recorded were 22630.7401 W/m²·K for CO2 and 23745.3600 W/m²·K for air, both occurring at a suction pressure of 49.9 bar The CHTC consistently increased in steps of approximately 1×10⁴ and 2×10⁴ W/m²·K Conversely, the lowest CHTC values for both scenarios were found at the highest suction pressure of 50.1 bar.
CO2-based CHTC Air blow-based CHTC
Figure 4.4 CO2-based, air-based CHTC and suction pressure correlation
A decrease in suction pressure significantly impacts the overall cooling load, as shown in Figure 4.4 The highest overall cooling load was observed at the lowest suction pressure of 49.9 bar, similar to the trend seen in CHTC Both sensible heat and latent heat increased continuously, with peak values of 1.769 kW and 0.3532 kW, respectively, occurring at the lowest suction pressure, inversely related to the reduction in suction pressure.
Figure 4.5 Cooling load and suction pressure correlation
Correlation between suction pressure, COP and overall cooling load (in CO2 terms) proved the enormous indirect affection of suction pressure on system COP as well as cooling load (Figure 4.5).
Table 4.4 Other quantities calculation database
Decreasing suction pressure by 1 bar over time consistently increases the coefficient of performance (COP) and cooling load, highlighting the significant impact of suction pressure on system efficiency As indicated by equation (4.17), the overall cooling load directly influences the COP, meaning that an increase in cooling load results in a higher COP However, an instability in the COP line pattern can occur even when the cooling load line appears stable, potentially due to ambient factors or measurement errors during experimentation.
COP CO2-based overall cooling load
C oo li n g lo ad ( kW ) C O P
Figure 4.6 COP, overall cooling load and suction pressure correlation
Numerical simulation results analysis
The 16% deviation observed in the evaporator outlet temperature, as shown in Table 4.5 and Figure 4.7, may be attributed to measurement errors or variations in the material properties of the geometrical objects Nonetheless, this deviation is considered acceptable due to its relatively low magnitude.
Table 4.5 Comparison between experiment and simulation for evaporator temperatures
Figure 4.7 Simulation result in temperature (left) and pressure (right) terms
The simulation results in this study indicate that pressure loss at the evaporator outlet is minimal, as evidenced by the data in Table 4.6 and Figure 4.7 With an approximate value of +0.047, the pressure loss can be considered negligible for the purposes of this research.
Table 4.6 Comparison between experiment and simulation for evaporator pressures
Figure 4.8 Simulation result in phase change terms (left) and thermographic image of the evaporator backside (right)
4.4.3 In terms of phase change
In theoretical estimations with minimal error, the vapor quality of the refrigerant at the evaporator outlet can increase by up to 0.05 for each 1 cm or longer section of the object, as observed in a 10 cm section, where the actual result is 0.059 (refer to Table 4.7 and Figure 4.8) Therefore, the simulation results are deemed acceptable for approximating an 18% calculation in vapor quality.
Table 4.7 Comparison between experiment and simulation for evaporator phase changes
CONCLUSION AND RECOMMENDATION
Conclusion
Analysis has provided following general conclusions:
+ Highest values of COP were obtained at valve inlet pressure of 78 bar in this study (4.52).
+ Minor decrease in suction pressure (from 50.1 to 49.9 bar) also resulted in decrease in room temperature values (from 29 o C to 27.7 o C).
Significant increases in the cooling heat transfer coefficient (CHTC) were recorded, with CO2 rising from 17,436.74 to 23,745.36 W/m²·K and air blow increasing from 16,632.42 to 22,630.74 W/m²·K, alongside a slight decrease in suction pressure from 50.1 to 49.9 bar This continuous rise in overall cooling load, with CO2 escalating from 1.9191 to 2.1227 kW and air blow from 1.9185 to 2.1222 kW, contributed to an improvement in the system's coefficient of performance (COP).
Recommendation
Intercooler and micro gas cooler installations are expected in order to improve system’s performance.
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The article by John R Thome and Gherhardt Ribatski provides a comprehensive overview of the latest advancements in two-phase flow, focusing on flow boiling heat transfer and pressure drop of CO2 in both macro- and microchannels Published in the International Journal of Refrigeration, this study highlights the critical factors influencing heat transfer efficiency and pressure dynamics in various channel sizes, contributing valuable insights for the refrigeration industry.
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The research conducted by J Yang, B Yu, and J Chen focuses on enhancing the predictive capabilities of a CO2 microchannel gas-cooler used in automobile air conditioning systems The study utilizes an improved genetic algorithm to analyze experimental data, contributing to advancements in refrigeration technology This work, published in the International Journal of Refrigeration, emphasizes the significance of accurate modeling in optimizing automotive cooling systems.
Nghiên cứu của ThS Nguyễn Trọng Hiếu, PGS.TS Đặng Thành Trung, ThS Lê Bá Tân, NCS Đoàn Minh Hùng và KS Nguyễn Hoàng Tuấn tập trung vào các đặc tính truyền nhiệt trong thiết bị bay hơi kênh micro sử dụng môi chất CO2 Nghiên cứu này được thực hiện thông qua phương pháp mô phỏng số và được trình bày trong Kỷ yếu hội nghị khoa học và công nghệ toàn quốc về cơ khí lần thứ IV.
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PGS TS Đặng Thành Trung đã nghiên cứu và phát triển hệ thống điều hòa không khí sử dụng thiết bị bay hơi kênh mini kết hợp với môi chất lạnh CO2 Mục tiêu của đề tài nghiên cứu khoa học và công nghệ cấp bộ này, được thực hiện vào tháng 3 năm 2018, là nhằm tiết kiệm năng lượng và bảo vệ môi trường.
PGS TS Đặng Thành Trung đã thực hiện nghiên cứu thực nghiệm về các thông số nhiệt động trong quá trình giãn nở và tiết lưu của hệ thống điều hòa không khí sử dụng môi chất CO2 Đề tài này thuộc chương trình KH&CN cấp trường trọng điểm năm 2017 tại Trường Đại học Sư phạm Kỹ thuật Tp.HCM, hoàn thành vào tháng 03/2017.
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Table A R744 - CO2 saturation properties (pressure) p t v (m 3 /kg) h (kJ/kg) s (kJ/kgK)
Table B Superheated and transcritical vapor properties of R744-CO2
(t boiling = 5.3 o C) t o C v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK tbh 0.01903
(t boiling = 28.68 o C) t o C v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK tbh 0.00638
8 MPa 9 MPa 10 MPa t o C v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK
11 MPa 12 MPa 13 MPa t o C v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK
14 MPa 15 MPa 16 MPa t o C v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK
17 MPa 18 MPa 19 MPa t o C v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK v m 3 /kg h kJ/kg s kJ/kgK