The most important substitute of the refrigerant is the carbon (IV) oxide. The gas is favorable when compared to the gases in terms of their environmental properties. This gas has low energy efficiency when applied in the air conditioner hence achallange.in the expansion process there is a major los known as the throttling loss. The expansion device, expander can be used in the recover the availability that was lost during the production. When comparing the basic carbon dioxide cycle with the current conventional system is more efficient since the estimation ranges from (40-60) % when subjected to the temperature range of (27-50) digress Celsius (Wilson & Korakianitis, 2014).
The main objective of the project is to determine the overall performance of the gas expander when applied in the residential homes air conditioning. There are several approaches that need to be followed ranging from the performance to feasibility standpoints. For a realistic system performance and evaluation to be gotten there should be proper selection of the expander to be used (Kalina, 2011).
The energy efficiency approximation of the system of air conditioning with the conventional air expansion valve is measured to be (4-6) % in the 28 degrees Celsius outdoor temperature. The ideal expander of the carbon (IV) oxide performance is approximated to be 25-38% efficient in the 28 degrees Celsius outdoor temperature. When comparing the ideal expander with the R22 system at temperature of 50 degrees Celsius, the expander is more efficient at 4-7% but when the efficiency of the expander is 80% then the carbon dioxide systems performance is better that that of the expander since its efficiency ranges from 16-17% at an outdoor temperature of 50degrees Celsius. When the exchange of the system valves are done by replacement with expander then the performance ranges from 19-31%. When the expanders efficiency of is 80%, the performance of the co2 is 10-21% at the temperature of 28 degree Celsius. But with the expander efficiency at 60% the systems efficiency of the energy ranges from 4-14% at the same temperature. The table shows the performance coefficient of the system (White, Parks & Markides, 2013).
(Al-Sarkhi, Jaber & Probert, 2016)
The expansion mechanism system has been analyzed by the following three positive displacements, the reciprocating piston, rotary piston and scroll.
The figure below show the summary of the performance obtained from the device as compressor and expander respectively (Austin & Sumathy, 2011).
There are internal actuated valves in the expanders that reciprocate and rotate. The valves control limits the frequency of the expanders to 1Hz. For the volume ration to be adjusted there is requirement for the control valves. The volume ratio internally remains constant when the there is no control on the inlet valve of the expander. The expender efficiency can be reduced when the ideal volume ratio imposed by the conditions of the operation, does not match the internal volume ratio. On conditions operation in a wide range there is change in optimized pressure to less than 25%, this is indicated by the performance simulation. At time the expander valves may not be necessary especially when the expander’s performance degrades due to over or under expansion. The efficiency of the expander remains constant at all conditions of operation when there is a minimal change in the ratio of pressure of the operation conditions over a wide range. The efficiency is 91% and 87% for the reciprocating piston compressor and that of the reciprocating piston expander respectively (Ma, Li & Tian, 2013).
The improvement of the energy efficiency from range of 41-81% can be done by the ideal gas expander when subjected to the temperature of about (28-50) degrees Celsius respectively. The performance of the system can be improved at range of (31-56) % when the efficiency of the expander is 80% and 60% when the efficiency is increased by the 22-38 %.there are no benefits from the line heat exchanger by the expander during the most operation condition. Due the increase in the capacity of the cooling capacity, there is reduction in the work output of the expander where the compressors input are increased (Lontsi, Hamandjoda, Fozao, Stouffs & Nganhou, 2013).
Dependability of the expander’s performance lies with the irreversible process during the expansions process. Positive displacement machine contains the following;
The increase in the heat coefficient by the heat transfer internally, can greatly affect the expanders performance due increase in evaporation. When there is zero net transfer between the fluid and the expander, expander performance can be affected 3%.when the expander dominate the heat transfer, it implies that operation is adiabatically and the net heat transfer is equivalent to zero. The measurement of the performance depends on the losses that are mentioned above. The table below shows the performance that is expected of the expanders and the compressors (Yanagisawa, Fukuta, Ogi & Hikichi, 2011).
The piston compressors performance is similar over the range of ratio of pressure. When there is high pressure, there is fall of the performance ratio behind the reciprocating piston compressors. The leakage a cross the tip of the scroll wrap dominates mechanisms in the expander. The performance of the scroll can be similar to the other mechanism if the leakage could be reduced. Special attention should put on the two valves operation due to impossibility to reed valve in the application of the expander. The losses range to about 10% when there is optimization of the volume ratio for one condition of operation but no adjustment is done. Some of the losses that were experienced by the machine were friction loss that depends on the design of the machine.
This report tries to determine the overall performance of the gas expander when applied in the residential homes air conditioning. There are several approaches that are taken into place to determine the optimum performance of the gas expander. Some of the techniques that are used are the approximation of the energy efficiency of the system of air conditioning with the conventional air expansion valve. During the analysis of the expander optimization results, it must be taken into consideration that a simulation of the devices was done with an ideal valve actuator. The valves used were actuated with pressurized gas while solenoid valves were used in reciprocating CO2?piston expanders. In both of the above projects, valve timing limited the expander speed to about 60 rpm. . For the full exploration of the mechanism of the expander and scroll compressor in a CO?2 system, it is necessary to have experimental investigations together with accurate data regarding the effective leakage gap size. The advantage here is that the valve control depends only on the speed of the shaft and not on external means. This project shows that there is exhibition of the problem of the ling valve by the first prototype.
The purpose of this project is to achieve the overall performance of the gas expander when applied in the residential homes air conditioning. This project tries to explain the operation of the expander thereby determining the performance of the expander. The project also tries to determine the how the expansion device, expander can be used to the recover the availability that was lost during the operation, to achieve a maximum efficiency, the expander mechanical work must be utilized directly. For the purposes of reduction of storage and conversion losses, there is need for the expander’s integration and the compressor on the shaft that are common for the work expander to supply some of the compressor shaft work. Simulations of the system has also0 been used to show that there can be an operation on the expander system close to or at the optimum high side pressure without extra control valves or variable expander displacement if the expander and the compressor displacement volumes are appropriately designed.
The rotary and reciprocating piston expanders require valves that are actuated externally. With valves controlled and activated externally, current and latest technologies restrict the frequency of the expander to order values of 1Hz. However, the scroll mechanism is not affected by the mentioned limitation. All mechanisms to adjust the volume ratio require some controls on the valve. The internal volume ratio of the expander remains constant if its inlet valves are not regulated. The expander frequency is reduced by under- or over-expansion if the ideal volume ratio established by the conditions of operation does not match the internal volume ratio. Simulation of the performance shows the volume ratio that is ideal in a carbon (IV) oxide optimized high pressure system over a broad range of operating conditions varies with less than one quarter. The degradation of the performance of the expander as a result of under- or over-expansion is less than 10%, showing that the expander valve control is no longer needed. As a result of the minor variation of ratio of the pressure over a broad range of conditions of operation in a system of CO?2, the expander as well as the compressor indicated efficiencies are? estimated to be constant at all the conditions of operation (Boyce, 2011). Efficiencies of a reciprocating piston expander that has been indicated and a reciprocating piston compressor with a size gap leakage of 10m are estimated at 87% and 91%, in that order. The effective isentropic efficiencies are 72% and 76% for the expander and the compressor, in that order. This is 15% less than the indicated efficiencies. At these effective isentropic frequencies, the COP of an optimized CO2 ? system of the expander is estimated at 2.4 at 50o?C and 6.6 at 28?o?C outdoor atmospheric temperature. In comparison to type of system known as conventional R22, the system of CO?2 expander exhibits a worse performance of 14% at high temperature conditions and operates 20% better at conditions of low temperature. It is estimated that at about 34?o?C, both the two systems exhibit same energy efficiency in their performance/operation.
To achieve a maximum efficiency, the expander mechanical work must be utilized directly. For the purposes of reduction of storage and conversion losses, there is need for integration of the expander and the compressor on the shaft that is common for work of the expander to supply some of the shaft work of the compressor. However, with this design, it is challenging to optimize the high side pressure, due to the fact that high side pressure lacks independence from the speed of the compressor, unless the expander has a volume with the variable displacement. There has been an analysis of a number of control options. Simulations of the system show that there can be an operation on the expander system close to or at the optimum high side pressure without extra control valves or variable expander displacement if the expander and the compressor displacement volumes are appropriately designed.
The figure below illustrates the difference between a cycle with isentropic and isenthalpic expansion in a CO?2 p-h diagram. The points 1, 2, 3, 4, are the cycle with isentropic operation of the expansion.
Difference in enthalpy, ?h?4’? h?4? ?is obtained in the nature of expansion work for the process of expansion to lead a constant entropy line. This is an implication that isentropic expansion indicated by the cycle, must possess a device for work extraction. The net input of work to the cycle can decrease by the work generated. At the same time, the cycle capacity is raised by the difference h?4’? h?4. ?Hence, the isentropic expansion raises the capacity and reduces the cycle work input (Qin, Chen, Sun & Wu, 2013).
The isentropic process of expansion is impossible technically because of leakage, friction and other sources of irreversibility that are unavoidable. In practical applications, a work extracting device experience a polytrophic process, leading to an inlet state of evaporator between points 4’ and 4 as illustrated in the above diagram (Lemort, Declaye & Quoilin, 2012).
The comparison of the practical process of the expanders and the isentropic imaginary process are characterized by the expander’s isentropic efficiency.
In similarity with the compressor isentropic efficiency, this statement is applicable only if it is adiabatic process. An isenthalpic throttling can be regarded as an expander having zero efficiency isentropically (He et al, 2013).
A first law analysis gives room to estimate the system performance improvement potential using the expander. The coordination of the analysis has been done by both a s system known as subcritical R22 and a trans critical CO2 system. An assumption is made that all the work from the process of expansion is recovered thus contributing to the total work input of the system (Sahoo, Sahoo & Saha, 2012).
The graph below gives a graphical simulation results over a wide range of ambient temperatures between 28?degrees Celsius to 50?degrees Celsius and expander efficiency between 0% and 100 %. In the absence of the expander, the R22 system operates equal to or better than the CO2 system at all ambient? temperatures, as shown in diagram 2a. The system baseline of R22 with expansion valve that are conventional underperforms when compared with a CO2? system having an isentropic expansion over considered ambient temperatures range. With both cycles having an isentropic expander, the CO?2 system outshines the R22 system at ambient temperatures of less than 38?o ?C, as illustrated in figure 2b (Badami & Mura, 2016).
For the estimation of expander and compressor performances, a universal simulation has been developed for positive displacement machines. The algorithm computes the device performance taking into account the heat transfer internally, internal leakage, and ratio of the volume that is unmatched and valve losses. The inputs of the program are the conditions of the boundary for which expansion and/or compression occur and the device with the positive displacement geometric data. The thermodynamic state gives the simulation outputs in the process, gas forces, work, and mass flow rate, heat transfer to and fro the fluid as well as the moments acting on the drive shafts and moving surfaces. The algorithms implementation that is performed on MATLAB Release 12 and uses an interface to Refprop 7 known as CEEE Props to compute the fluid properties. For simulation of single operating points, for batch operation the program has a provision for graphical user interface as well as an interface to Microsoft Excel (Kuriyama et al, 2016).
Fluid properties
In a control volume, the properties of the fluid are bulk properties (average). The assumption is that during each time step the properties are constant. There is a calculation of the properties for a thermodynamic equilibrium. Therefore, the simulation of the process is carried out a s a quasi-steady-state process (Galindo et al, 2015).
Valves
During the modeling of the valves, it is assumed that all the valves used are ideal. Similarly, It is assumed that there is full opening of reed (pressure actuated) valves without any time delay. In addition, it is assumed that position actuated valves such as scroll expanders close and open fully instantaneously. The flow across the valves is simulated as circular orifice flow (Branchini, De Pascale & Peretto, 2013).
All expander valves are simulated as position actuated valves, and the use of reed valves is impossible in expanders. For the responds of the valve with respect to reservoir and the pressures of the pocket, it is necessary to incorporate an external control. However it is a challenge to build a valve controlled externally that operates very fast and has an acceptable flow resistance for AC applications in buildings for residential. All expanders possess built-in volume ratios since they have the discharge and suction timing tied to the angle of crankshaft. If the ideal volume ratio that is built-in is not equal to the real ratio of volume of the expander imposed by the conditions of the boundary, efficiency that is indicated reduces due to under /over expansion. Figure 5 below illustrates the effect of under- or over-expansion (Radebaugh, 2013).
Oil
Oil presence in the volume of the control is ignored, unless there are impacts that are significant on the leakage of the refrigerant. To analyze the expander and compressor mechanisms in this paper, only compressor with the rotary piston model used the oil effect (Clemente, Micheli, Reini & Taccani, 2013).
Internal leakage flow
There are several effects of the internal leakages on the expanders and the compressor. Most of the research that has been done shows that leakage flow depends on the internal pressure difference. The computation of the instantaneous leakage flow rate is done by determining the speed of sound. For lunar to turbulent flow to be achieved the following computation can be done.
The diameter that is effective is estimated as 0.666 of the diameter of the hydraulics. Factor of friction for the flow of the turbulent is ¾.the mine losses has the following losses ½.
Leak paths are categorized into two different categories that is, the leak of the tip and leak of the flank. For the tip, the flow is consider of the minimal height and large width, while in the case of the flank, the passage is basically found in the devices such as the scroll.
There is improvement of the sealing during operation by the use of the oil in the cases of refrigerants. But when the pressure inside become high, the refrigerators state to leak due to the dissolution of the air by the high pressure (Moss, Roskilly & Nanda, 2015).
Expansion process can be done by two simulation methods
The integration of the numerical transient
Find the solutions of the numerical leakage-flow rate, thereby helping in the energy and mass conservation (Smith, Stoši? & Aldis, 2015).
Solution of the problem by the numerical integration
The use of the runge kutta theorem,there is observation of the mass and energy conservatiobn violation by the leakage flow.the effect of the valve leak flow can onmly be fealt in the control volume that is cuurently experienced and the effect of the lea flow is felt in all the control volumes.after tha formation of the control volume in the process of suction,the connection of the leak and the pocket closer to thecenter of the scroll is done.thei may lead to the incraese in presuer as some of the amount may laek tyo the control voluume.maintaince of the volume may be at the point in which high pressuere pocjket is when there is single rotation of the crankshaft.there is connection of the control vvolume with the walls of the scroll.the may lead to reduction of the flow rate due to the pressuere difference.furtehrmore there is increase in the mass that is contained in the control volume andf the pressure pocket hence leading to the violation of the maass and energy consevesion (Bao & Zhao, 2013).
Simultanious computation of the leak flow can be used to avoid this kind of ua problem.the production of the mass and the internal energy is done by the use iof the intergartion starting from the earler gueses of the for the leak flow.the computation of the leak flow is based on the pressure profile.to achive stady operation of the exanders,the should be no change in the applied presssure of the leak flow rate.hence similar pressuere must be peoduced during the c omputation as shown below.
But F is the function of the system that contains the massand the enregy of the expander (Duc, Chauvy & Herri, 2017).
The modeling procedure covers several steps when dealing with the algorithms. The approximation of the irreversible heat flow process is considered during the process. The volume of the two control systems undergoes compression and immerged to form one control volume at the scrolls Centre (Vandor & Greenberg, 2017).
Expander simulation commences at the intake process. When the intake process begins, the angle of crankshaft rotation is zero. Gas starts flowing into the pocket when the shaft turns, increasing the angle of crankshaft. At a given angle, the pocket disconnects from the intake reservoir, expanding the gas. On reaching the angle of discharge the shaft opens the pocket to the low-pressure reservoir, expelling the gas from the chamber.
The angle of crankshaft leading to completion of one cycle relies on the expander type. For a reciprocating piston expander, for example, at crankshaft angle 2?o and 4?o ? for a rotary piston. On the other hand, the angle relies on wraps lengths for a scroll expander (Boretti, 2012).
Temperatures of the fluid, wall surface area and heat transfer coefficient between fluid and wall affect the transfer of heat between the walls of the control volume the fluid contained in a control volume. The device geometry gives the surface area of the device. The temperature of the wall is taken as the conditions of the boundary while the solution of the fluid is found. The coefficient of heat transfer relies on the fluid movement and the fluid state in the control volume. The movement of the fluid fluctuates with time and relies on the control volume geometry strongly (Wang, Zhang, Peng & Shu, 2011).
There has been little attention paid to literature of positive displacement expander’s internal heat transfer. Modified correlations for the coefficient of transfer of heat in compressors used in the absence of experimental correlations for the transfer heat’s coefficient. Besides convection, evaporation of liquid refrigerant at the pocket wall enhances expander’s heat transfer. The coefficient of heat transfer in turbulent pipe flow is typically by factor 5 in the presence of evaporation. Thus, the coefficients of transfer of heat acquired with equation 33 are multiplied by a factor of 5 for simulations of expander (Vandor, D., & Greenberg, 2018).
An analysis on the performance of the following expander and compressor mechanisms has been done:
For the sake of fairness in comparison, individual mechanism geometry is optimized to be used as an expander and a compressor under constant boundary conditions). The chosen values are for 10kW cycle of CO?2 expander at the conditions of the AHRAE A, assuming the expander and the efficiency of the compressor isentropic is at 75%. The cooler of the gas approach temperature is in the order of 3K while the evaporating temperature is approximated at 7?o?C. It is important to note that the cycle excludes an SLHX. The devices’ valve timing, having a built-in volume ratio (all scroll compressors and expanders), is optimized for each of the operating conditions (Berry & Pruitt, 2016).
|
Expander |
Compressor |
|
|
P?inlet[? MPa] |
8.564 |
4.125 |
|
?inlet[? kgm?-3?] |
524.3 |
120.1 |
|
P?outlet[? MPa] |
4.175 |
8.675 |
|
Transfer of Heat |
0 |
0 |
The volume of the displacement of the compressor is 18.46cm?3 ?while the expander’s volume displacement is about 7.93cm3? ?at 3300 rpm for each of the devices (Bell & Partridge, 2013).
The assumption here is that there are two cylinders that are possessed by the reciprocating piston expander and compressor. The estimation of the clearance volume is done by the gap between the piston and head of the cylinder at 1mm TDC. With a fixed volume of displacement, the cylinder geometry is a one function variable, for example the cylinder to height diameter ratio (Glavatskaya, Podevin, Lemort, Shonda & Descombes, 2012).
Similarly, just like the cylinder aspect ratio, the inlet and the outlet ports cross-sectional area also affects the performance. In assumption, a particular fraction of bore area serves as the valve area. It is estimated that the valve area connecting to the high pressure reservoir is two-third of the low pressure valve area (Gord & Jannatabadi, 2014).
The piston rings provide sealing. It is assumed that the piston rings flow resistance is equal to a flow passage of length 3 mm and height 10m per unit width. The low side pressure is given by the shell pressure in both the expander and the compressor simulation (Coney, Abdallah & Richards, 2014).
The figure illustrates the volumetric and indicated efficiencies of a compressor with the reciprocating piston using the aspect ratios in wide range. The valve areas are indicated by the graphs and also show the performance effect. There is sharply increase in the indicated efficiency from smaller magnitude aspect ratios, goes through peak 0.5 aspect ratios and slowly reduces at ratios higher than 0.5 to 1. Valve losses effects and leakage of performance are directly reflected on this curve. The available area for valves is small at small aspect ratios so that the valve losses dominate the process. Increase in the bore area renders the valve losses insignificant, but the length of sealing on the piston rises leading to leakage rates of flow that is higher.
The effective port area of a practical expander is in the order of 50% of the bore area, estimation aspect ratio that is optimum at 1 in this design, the resultant efficiency that is indicated becomes 91% while the volumetric efficiency is at 84% and each cylinder bore and stroke is estimated at 2.3cm (Viteri, 2014).
A range of aspect ratios have been used to find the indicated efficiency of a reciprocating piston (figure above). The curves illustrate the impact of valve area on performance. The absence of volumetric efficiency is due to its inability of to be used in characterization of performance. With an assumption that the port area of a practical expander is in the order of 50% of the bore area, the aspect ratio that is optimum is estimated at 1, the resultant efficiency that is indicated becomes 88% and each bore of the cylinder and stroke of the cylinder is estimated at 1.7cm (Abdelmalek, 2015).
This project requires a lot in order to be completed successfully. Some of the requirements are the gas expander and the compressors, laboratory for the experiment and the capital for the purchase of the equipment.
Discussion and Conclusion
At gap sizes of higher leakage, this mechanism exhibits a strong drawback in comparison to others due to the low scroll compressor and the expander performance. This strong weakness is not evident in conventional refrigerants scroll compressors, because of smaller differences in pressure compared to CO?2 and good sealing techniques. For the full exploration of the mechanism of the expander and scroll compressor in a CO?2 system, it is necessary to have experimental investigations together with accurate data regarding the effective leakage gap size (Wang, Wu, Ma, Liu & Yu, 2014).
During the analysis of the expander optimization results, it must be taken into consideration that a simulation of the devices was done with an ideal valve actuator. For the scroll expander, the assumption may not be critical, since closing and opening of the devices’ ports relies purely on the angle of crankshaft. However, in the expander’s with the rotary and reciprocating piston, the expansion’s compartment is permanently connected to the outlet port and inlet port. Thus, there must be an external control of the port timing. Both the controlling of the outlet and the inlet ports must be done for the piston that is reciprocating (Habets & Kimmel, 1998). The valves used were actuated with pressurized gas while solenoid valves were used in reciprocating CO2?piston expanders. In both of the above projects, valve timing limited the expander speed to about 60 rpm. Examination of a rotating disc control mechanism being slots, obtaining it from hydraulic applications. The advantage here is that the control valve depends on shaft’s speed. This project shows that the prototype that first is exhibited by the valve leakage problems (R., & Weerasinghe, 2016).
For the case of the rotary piston expander, only the inlet port requires control. By operating at the same shaft speed, the rotary piston expander inlet shaft opens twice as long as the expander with the reciprocating piston inlet port, leading to a potential reduction in the problems associated with valves controlled externally. Similarly, the adoption of the rotating disc concept of a as rotary piston expander control valve may be possible (Kucerija, 2014).
As shown in figure below, the variation of the efficiency is that indicated of the expander over the ratio of pressure shows that it is unnecessary to the conditions of operation the ratio volume that is built-in over the considered range. The ratio of the volume changes from 2.3 to 2.9 for 32 ?o?C and 55?o?C ambient conditions. For a mismatch of volume ratio on the order of 25%, figure 5 helps to estimate the performance degradation. There is a 5% drop of the indicated efficiency (Stobart & Weerasinghe, 2016).
The trapped gas in the rotary clearance and the compressors with the reciprocating piston goes through re-expansion until the pressure of the pocket nears the line pressure suction. However, in the rotary and reciprocating piston expanders, the increase in pocket volume opens the inlet valve. The clearance volume gas is at low density as well as low pressure.
The incoming gas has high density and high pressure. Immediately after opening of the inlet valves, the gas in the suction line goes through abrupt, expansion that is irreversible until the level pocket pressure equals the high line pressure. This results to introduction of extra losses in the piston device expansion process. Theoretically, accurate valve timing enables the elimination of these losses. The scroll compressor outlet and inlet valves uncover at certain crankshaft angles without external actuators in place. However, without extra control input, operation of a scroll expander happens at a constant volume ratio, that necessarily does equal the optimum volume ratio brought about by the boundary conditions. A reed valve is usually used to adjust the volume of a scroll compressor after the discharge port, for the scroll’s center pocket to open to the reservoir only when pressure of the pocket matches the pressure of the reservoir (Konukman, & Akman, 2015).
On the other hand, the inlet port is the port at the scroll’s Centre, for a scroll expander. It is not possible to use a reed valve for timing here. It is necessary to have an external control for adjusting the volume ratio that is built to the conditions of operation (Park & Kaviany, 2012).
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