Page 116

Evaluation of Coefficient of Performance of a Vapour Compression Refrigeration System with Change in Length of Condenser Saragadam Vinay Kumar 1, Karteek Naidu Mariserla 2 #1Student of M.Tech in Mechanical Engineering Department with specialization in Thermal Engineering, #2 Assist.Prof, Department of Mechanical Engineering in Gokul group of Institutions, Bobbili, vizianagaram, AP. Abstract- The execution of heat transfer exchange is a standout amongst the most vital examination regions in the field of warm designing. There are an extensive number of refrigerants, which are utilized to exchange heat from low temperature supply to high temperature store by utilizing vapor pressure refrigeration framework. There are different obstructions confronted in working of diverse refrigerants because of their natural effect (R11, R12), harmfulness (NH3), c ombustibility (HC) and high weight (CO2); which makes them a bigger number of unsafe than other working liquids as indicated by security and ecological issues. Scientists watched the execution of diverse natural welldisposed refrigerants and their blends in distinctive extents. They additionally watched the impact of working parameters like measurements of slender tube, working weights and working temperatures, which influence the coefficient of execution (COP) of vapor pressure refrigeration framework. From the writing there is by all accounts need of new proficient, least an Earth-wide temperature boost potential (GWP), least ozone exhaustion potential (ODP) and natural neighborly refrigerants. The conduct of execution parameters of a basic vapor pressure refrigeration framework were concentrated on while it's working under transient conditions happened amid cooling of a settled mass of brackish water from beginning room temperature to below zero refrigeration temperature. The impacts of distinctive lengths of fine tube over these Characteristics have likewise been researched. It was inferred that with the Continually falling temperature over evaporator, refilling of it with an ever increasing amount Fluid refrigerant reasons multifold increment in warmth exchange coefficient which helps in keeping up refrigeration rate at falling temperature. In this test we go above and beyond and examine the impact of variety in condenser tube length. Bigger condenser tube diminishes the inclination of refilling of evaporator however offers less 'evaporator temperature' successful in lower scope of refrigeration temperature. Shorter condenser tube guarantees higher COP at first and increments to a given point. The condenser tube length must be enhanced for greatest general normal COP of the framework for the complete determined cooling occupation. Index Terms Vapour Compression Refrigeration System, Refrigerant, Cop, Odp, Gwp. I. Introduction: A fundamental vapor compression refrigeration framework comprises of four crucial segments, to be specific compressor, condenser, development valve and evaporator. The individual execution attributes of these segments have been examined in before addresses. In any case, in a real framework these parts work as one. The execution of a complete framework is a parity's aftereffect between these four parts. For instance, when the warmth sink temperature changes, it influences the condenser's execution, which thus, influences the extension's execution gadget, evaporator and the compressor. It is found in that development valve and compressor work in such a way, to the point that the mass stream rate through the two parts is the same at consistent state. The equalization point at consistent state was gotten by likening the mass stream rates through these segments. This is an illustration of adjusting two parts. Comparable methodology can be stretched out to incorporate the other two parts additionally, so an equalization point for the whole framework can be acquired by considering the individual attributes. On a basic level, the parity point for the framework can be acquired either by a graphical strategy or by an expository system. In graphical strategy, the execution of two reliant segments is plotted for the same two variables of regular hobby. For instance, mass stream rate and evaporator temperature (or weight) are plotted along y and x tomahawks separately for blend of compressor development gadget at steady condenser temperature. The purpose of crossing point of the two subsequent bends will show the conditions at which the mass stream rate and evaporator temperature will be same for the two segments. This point is known as the parity point and in consistent state the blend will accomplish these conditions. www.ijmca.org Page 116

In explanatory strategy, the mass stream rate through development valve can be spoken to by a logarithmic comparison as far as evaporator and condenser temperatures. Essentially, the mass stream rate through a given compressor can likewise be spoken to by an arithmetical mathematical statement as far as evaporator and condenser temperatures by relapse investigation of trial or logical information. The equalization purpose of the two segments can be gotten by synchronous arrangement of the two mathematical comparisons. Since the graphical strategy utilizes two-dimensional plots, it considers just two parts at once while the framework investigation by scientific means can consider more than two segments at the same time. Further, considering time variety of parameters in type of differential comparisons can mimic the dynamic execution too. Steady state framework investigation will include concurrent arrangement of mathematical comparisons. In this parity purposes of consolidating unit, compressor-evaporator mix have been considered for representation. As a first step the execution information of modern segments is exhibited as plots or comparisons. The crude information for this reason can be gotten from the lists of makers. These are plotted either specifically or in the wake of preparing as far as obliged variables. II. Reciprocating compressor performance characteristics: The power requirement and mass flow rate as function of evaporator temperature with condenser temperature as a parameter were presented on compressors. For the purpose of balancing, the refrigeration capacity is required as a function of evaporator and condenser temperatures. This can be easily determined by considering the refrigeration cycle or from the catalogue data of the manufacturer. Figure 25.1 shows a theoretical single stage saturated cycle on T-s chart. Fig: A single stage, saturated vapour compression refrigeration cycle For the above cycle, the refrigeration capacity and power input to compressor are given by: At a given condenser temperature the cooling capacity associated with mass flow rate given by a compressor increases as the evaporator temperature increases. On the other hand, for a given evaporator temperature, the cooling capacity decreases with increase in condenser temperature. These characteristics are shown graphically Capacity e T c= o C T c= o C T c= o C Evaporator e Fig1. Variation of refrigeration capacity of a reciprocating compressor with evaporator and condenser temperatures at a fixed RPM The following equation may represent the above trends: Qe=a1+a2Te+a3Te2+a4Tc+a5Tc2+a6TeTc+a7Te2 Tc+a8TeTc2+a9Te2Tc2 Where T e and T c are evaporator and condenser temperatures, respectively. The a 1 to a 9 are constants which can be determined by curve fitting the experimental or manufacturers data using least square method, or by solving nine simultaneous equations of the type (25.3) for the nine constants a i using nine values of Q e from given catalogue data for various values of T e and T c. Similar expression can be obtained for power input to the compressor. Ion can be obtained for power input to the compressor. II. Refrigeration System: Working Principle and Construction Refrigeration system is based upon the Clausius statement of second law of thermodynamics. This statement shows, It is impossible to construct a device which, operating in a cycle, will produce no affect other than the transfer of heat from a cooler to a hotter body. The construction of vapour compression refrigeration system is illustrated in figure 1. This system consists of four basic components, i.e. a compressor, an evaporator, a condenser and capillary tubes. Here the compressor delivery head, discharge line, condenser and liquid line form the high pressure side of the system. The expansion line, evaporator, suction line and compressor suction head form the low pressure side of the system. In plants with a large amount of refrigerant charge, a receiver is installed in the liquid line. A drier is also installed in the liquid line. The drier contains silica gel and absorb traces of moisture presented in the liquid refrigerants so that it does not www.ijmca.org Page 117

enter the narrow cross section of the expansion device causing moisture chocking by freezing. Fig 2: Schematic of the Vapour Compression Refrigeration System [5]. B. Processes Involved in Vapour Compression Refrigeration System: Fig 3: Pressure-Enthalpy Graph for Vapour Compression Refrigeration System [11]. Process 1 2: Isentropic compression in compressor. Process 2 3: Constant pressure heat rejection in condenser. Process 3 4: Isenthalpic expansion in expansion device. Process 4 1: Constant pressure heat absorption in evaporator. C. Refrigerant: The working fluid used to transfer the heat from low temperature reservoir to high temperature reservoir is called refrigerant. There are different types of refrigerant which are described as followings. CFC: They are molecules composed of carbon, chlorine and fluorine. They are stable, allowing them to reach the stratosphere without too many problems. It contributes to the destruction of the ozone layer. These are R11, R12, R113, R500, R502 etc. HCFC: They are molecules composed of carbon, chlorine, fluorine and hydrogen. They are less stable than CFCs, destroy ozone and to a lesser extent. These are R22, R123, R124, R401a etc. HFC: They are molecules composed of carbon, fluorine and hydrogen. They do not contain chlorine and therefore do not participate in the destruction of the ozone layer. This is known as substitution substance. Restrictions on this family of gas are currently limited. Within the European Union, the HFC will be banned from air conditioners for cars from 2011.These are R134a. 1. Mixture of refrigerants: They can be classified according to the type of fluorinated components they contain. They are also distinguished by the fact that some mixtures are: Zeotropic: in a state change (condensation, evaporation), the temperature varies. These are R404a, R407a and R410a etc. Azeotropes: they behave like pure, with no change in temperature during the change of state. These are R500, R502 and R507a etc. 2. Ammonia (NH 3 ) or R717 Fluid inorganic thermodynamically is an excellent refrigerant for evaporation temperatures between - 35 C to 2 C. But it is a fluid dangerous toxic and flammable, so it is generally used in industrial refrigeration. 3. Hydrocarbons (HC) as R290, R600a This is primarily propane (R290), butane (R600) and isobutene (R600a). These fluids have good thermodynamic properties, but are dangerous because of their flammability. The world of the cold has always been wary of these fluids, even if they have reappeared recently in refrigerators and insulating foams. Their future use in air conditioning seems unlikely, given the cost of setting both mechanical and electrical safety. 4. Carbon dioxide (CO 2 ) or R744 This is inorganic, non-toxic, nonflammable, but inefficient in thermodynamics. Its use would involve high pressure and special compressors. Currently, specialists in air conditioning and refrigeration are interested again by: Its low environmental impact (ODP = 0, GWP = 1); The low specific volume resulting in facilities with low volume (small leak); It has the distinction of having a low critical temperature at 31 C at a pressure of 73.6 bars. Refrigerant. There are different types of refrigerant which are described as followings. CFC: They are www.ijmca.org Page 118

molecules composed of carbon, chlorine and fluorine. They are stable, allowing them to reach the stratosphere without too many problems. It contributes to the destruction of the ozone layer. These are R11, R12, R113, R500, R502 etc. HCFC: They are molecules composed of carbon, chlorine, fluorine and hydrogen. They are less stable than CFCs, destroy ozone and to a lesser extent. These are R22, R123, R124, R401a etc. HFC: They are molecules composed of carbon, fluorine and hydrogen. They do not contain chlorine and therefore do not participate in the destruction of the ozone layer. This is known as substitution substance. Restrictions on this family of gas are currently limited. Within the European Union, the HFC will be banned from air conditioners for cars from 2011.These are R134a. 5. Mixture of refrigerants: They can be classified according to the type of fluorinated components they contain. They are also distinguished by the fact that some mixtures are: Zeotropic: in a state change (condensation, evaporation), the temperature varies. These are R404a, R407a and R410a etc. Azeotropes: they behave like pure, with no change in temperature during the change of state. These are R500, R502 and R507a etc. 6. Ammonia (NH 3 ) or R717 Fluid inorganic thermodynamically is an excellent refrigerant for evaporation temperatures between - 35 C to 2 C. But it is a fluid dangerous toxic and flammable, so it is generally used in industrial refrigeration. 7. Hydrocarbons (HC) as R290, R600a This is primarily propane (R290), butane (R600) and isobutene (R600a). These fluids have good thermodynamic properties, but are dangerous because of their flammability. The world of the cold has always been wary of these fluids, even if they have reappeared recently in refrigerators and insulating foams. Their future use in air conditioning seems unlikely, given the cost of setting both mechanical and electrical safety. 8. Carbon dioxide (CO 2 ) or R744 This is inorganic, non-toxic, non flammable, but inefficient in thermodynamics. Its use would involve high pressure and special compressors. Currently, specialists in air conditioning and refrigeration are interested again by: Its low environmental impact (ODP = 0, GWP = 1); The low specific volume resulting in facilities with low volume (small leak); It has the distinction of having a low critical temperature at 31 C at a pressure of 73.6 bars. III. Evaporator Performance Evaporator is also a heat exchanger just like condenser. For the sake of illustration, consider an evaporator that is used for chilling a brine. The cooling capacity of brine chiller is shown in Fig. 4 as a function of brine flow rate for different values of LMTD of evaporator. The brine side heat transfer coefficient hb increases as the brine flow rate increases as a result, the overall heat transfer coefficient of the evaporator increases. Figure 4 shows that the cooling capacity increases with flow rate for fixed LMTDe for this reason. Capacit e LMe= o C LMe= o C LMe= o C LMe= o C Brine flow Fig.4: Evaporator performance with brine flow rate and LMTDe One can obtain the data for cooling capacity at various brine inlet temperatures from the characteristics of evaporator as shown in Fig.4. For example, if a plot for brine inlet temperature Tb,i of 10oC is required, then we may choose an LMTDe of 5oC and read the capacity Qe for the chosen brine flow rate mb. Then the brine outlet temperature Tb,o is obtained from the equation: Qe = mbcpb (Tb,i - Tb,o) (25.11) Then the evaporator temperature Te is obtained from the expression for LMTDe: LMTDe = Tb,i Tb,o (25.12) ln Tb,i Te Tb,o Te The capacity Qe and evaporator temperature Te are determined for different values of LMTDe for a fixed brine flow rate and brine inlet temperature of 10oC. Figure 5 shows a plot obtained by this method. In this plot the brine flow rate is constant hence the www.ijmca.org Page 119

brine side heat transfer coefficient is constant. If the evaporation heat transfer coefficient was also constant then overall heat transfer coefficient will also be constant and these lines will be straight lines. The evaporation heat transfer coefficient increases with increases in evaporator temperature hence these lines deviate slightly from straight lines. The capacity for these lines may be expressed as follows: Qe = c0(tb,i-te)+c1(tb,i -Te)2 (25.13) Capacit e o C 1 o C Evaporator e Brine inlet b = o C Brine inletb = o C Brine inletb = o C 1 o C Fig.5: Performance characteristics of evaporator at fixed brine flow rate IV. Expansion valve Characteristics: The characteristics of expansion valve play an important role in deciding the conditions achieved by the refrigeration system. It was that compressor and expansion valve seek an evaporator temperature such that under steady state conditions, the mass flow rate is same through the compressor and expansion valve. This was the result under the constraint that the condenser and evaporator have sufficiently large heat transfer areas and do not influence the performance of expansion device and compressor. In this chapter it is assumed that the expansion valve is capable of providing sufficient mass flow rate at all condenser and evaporator temperatures. This is assumed to simplify the matching problem. A float type of expansion valve or thermostatic expansion valve will meet this requirement. If the analysis is being done by computational method then the valve performance may also be included with some additional computational effort. V. Condensing unit: As mentioned before, if graphical procedure is used to find performance evaluation of various components, then only two components can be considered at a time. In view of this the first subsystem considered is the condensing unit. Condensing unit is a combination of compressor and condenser. This unit draws refrigerant from the evaporator, compresses it in the compressor, condenses it in the condenser and then feeds the condensed liquid refrigerant to the expansion valve. It is available off-the-shelf as a packaged unit from the manufacturer with matched set of compressor, compressor motor and condenser along with reservoir and controls. This may be air-cooled or water-cooled unit which may be installed as an outdoor unit. The performance of condensing unit as function of evaporator temperature is obtained by combining the cooling capacity versus evaporator temperature characteristics of compressor and condenser. First we consider cooling capacity versus evaporator temperature assuming the compressor sped, the temperature and mass flow rate and entering water to condenser to be constant. This matching is obtained by superimposing the compressor performance curve given in Fig.25.2 on the condenser performance given in Fig.3 as shown in Fig.7. The intersection of compressor and condenser characteristics is at point A for 30oC condenser temperature. The combination of compressor and condenser will achieve a cooling capacity and evaporator temperature corresponding to this point at a condensing temperature of 30oC. Similarly, points B and C are the intersections at condenser temperatures of 35 and 40oC, respectively. These points are called balance points and the line A- B-C is called the performance characteristics of the condensing unit. Fig.7: Performance characteristics of a condensing unit as a function of evaporator and condensing temperatures It is observed that as the evaporator temperature decreases, the condensing temperature for the combination also decreases. This is explained as follows: at lower evaporator temperatures, the volumetric efficiency and the mass flow rate through the compressor decreases. This decreases the load on the condenser. A large condenser heat transfer area is available for small mass flow rate, hence condensation can occur at lower condenser temperature. It is also seen that as the evaporator temperature decreases, the refrigeration capacity of the condensing unit also decreases. This is due to the lower mass flow rate through the compressor due to lower volumetric efficiency and lower vapor density at compressor inlet. Figure 8 shows the variation of refrigeration capacity of the condensing unit with variation in inlet water temperature to the condenser. This is obtained by superimposition of compressor characteristics of Fig.2 on the variation of condenser performance with inlet water temperature given in Fig.4. The two figures are shown side-by-side. At constant evaporator temperature of say, 5oC and condenser temperature of 30oC, the inlet water temperature corresponding to point D is required to www.ijmca.org Page 120

match the two components. Points E and F are the balance points at condenser temperatures of 35 and 40oC respectively. Line DEF is the characteristics of the condensing unit at an evaporator temperature of 5oC. It is observed that the cooling capacity decreases as the inlet water temperature to condenser increases. Fig.6: Performance of the condensing unit as a function of water temperature at condenser inlet These characteristics can also be obtained by simultaneous solution of Eqns. (25.3) and (25.9) for constant water temperature at condenser inlet and constant water flow rate. For example, we wish to find the condenser temperature and capacity for a given evaporator temperature of say 10oC. An iterative procedure may be devised as follows: VII. Performance of complete system - condensing unit and evaporator: In steady state, a balance condition must prevail between all the components, that is, between condensing unit and evaporator assuming that the expansion valve will provide appropriate mass flow rate. This confluence will represent the performance of complete single-stage vapour compression refrigeration system. The combined curves will also give insight into the offdesign performance of the system and operational problems. Superimposing Fig.6 for the evaporator characteristics and Fig.7 for condensing unit characteristics yields the balance point of the system. This is shown in Fig.7. The characteristic curve shown as for constant water temperature at condenser inlet, constant flow rate to the condenser, constant compressor speed and constant brine temperature at the inlet to the evaporator. The point of intersection of the two curves gives the refrigeration capacity and the evaporator Capacit e y, Q 5 o C A B C 5 o C Evaporato (Brine1 o C) inlet = 0 Condensing 1 o C 5 Excess capacity (i) For Te = 10oC assume a condensing temperature Tc = 35oC (ii) Find Qe from Eqn.(25.3) (iii) Substitution of Te = 10oC and Qe in Eqn.(25.9) will yield a quadratic equation for Tc. The value of Tc is found and checked against the assumed value of Tc (35oC being the first iterate) and iteration is continued until the calculated value matches with the assumed value of condenser temperature. Evaporator e Fig.7: Performance of the complete system as an temperature, T intersection of evaporator and condensing unit characteristics at a brine inlet temperature of 10oC temperature that the system will achieve. One can study the response of the system in transient state also by this figure. In a transient state, say the evaporator temperature is 5oC. The figure shows that at this point the condensing unit has a capacity corresponding to point B while the evaporator has capacity corresponding to a lower value at C. Hence the condensing unit has excess capacity. The excess capacity will reduce the temperature of refrigerant and the metallic wall of the evaporator. This will continue until the balance point of 3oC is reached at point A. Figure 8 shows the effect of brine mass flow rate compared to that at the balance point. If the brine flow rate is increased, it is observed that cooling www.ijmca.org Page 121

capacity increases to point D. At higher flow rate the overall heat transfer coefficient increases while (Tb,i- Tb,o) decreases permitting a larger mean temperature difference between refrigerant and brine. Therefore with increase in mass flow rate of brine, the cooling capacity increases. The pump power also increases for the increased brine mass flow rate. Hence one has to make a compromise between increased capacity and increased cost of pump power. Figure 25.10 shows the condition for lower brine flow rate when the heat transfer coefficient on brine side decreases and temperature difference (Tb,i -Tb,o) increases. This is referred to as starving of evaporator. Capacity e, Q A D Condensing unit Increased brine flow rate Design flow rate Evaporator starving Evaporator e temperature,t Fig.8: Influence of brine flow rate on system cooling capacity VIII. Effect of expansion valve: So far we have considered the balance between compressor, condenser and evaporator assuming that expansion valve can feed sufficient refrigerant to the evaporator so that heat transfer surface of the evaporator is wetted with refrigerant. Thermostatic expansion valve meets this requirement. Automatic expansion valve and capillary tube as observed in Chapter 24, result in a condition where sufficient quantity of refrigerant could not be supplied to evaporator. This condition was referred to as starving of evaporator. Starving reduces the heat transfer coefficient in evaporator since there is not sufficient refrigerant to wet the heat transfer surface consequently the cooling capacity reduces. There are other conditions also which may lead to this situation. These are as follows: (i) Expansion valve is too small, (ii) Some vapour is present in the liquid entering the expansion valve, and (iii) Pressure difference across the expansion valve is small If the refrigerant charge in the system is small then condition (ii) is likely to occur. Also if the frictional pressure drop in the liquid line is large or the valve is located at higher elevation than condenser then this condition may occur. During winter months the ambient temperature is low hence in air-cooled condenser the condenser pressure is low and the difference between evaporator and condenser pressure is small, as a result the starving condition (iii) is likely to occur. In this condition the expansion valve does not feed sufficient refrigerant to the evaporator since the driving force; the pressure difference across the expansion valve is small. The evaporator pressure also decreases in response to drop in condenser pressure. The evaporator pressure may become so low that mass flow rate through compressor may decrease due to lower volumetric efficiency. Hermetic compressor depends upon the mass flow rate of refrigerant for cooling on motor and compressor. This may be adversely affected under starved condition. IX. Conclusion: The methods presented in this are useful when compressor, condenser, evaporator and expansion valve have been selected and the performance of combined system is desired. This analysis may not be useful in selecting the initial equipment. The techniques presented in this chapter are useful in predicting system performance for off-design conditions like a change in ambient temperature, condenser inlet water temperature and brine inlet temperature etc. The power requirement of the compressor has not been given due emphasis in the analysis. In fact, an equation similar to Eqn. (25.3) may be written for this also. This can also be found from known values of condenser and evaporator loads. An important aspect of refrigeration system performance is the sensitivity analysis which deals with % change in, say cooling capacity with % change in capacity of individual components like the compressor size, heat transfer area of evaporator and condenser etc. This can easily be done by mathematical simulation using the performance characteristics of the components given by empirical equations. It has been shown in Stoecker and Jones that compressor capacity has the dominant effect on system capacity and evaporator is next in importance. An increase in compressor capacity by 10% has the effect of 6.3% increase in system capacity. A 10% www.ijmca.org Page 122

increase in evaporator gives 2.1% increase in system capacity, while 10% increase in condenser gives 1.3 % increase in system capacity. Such a data along with the relative costs of the components can be used for optimization of the first cost of the system. References [1] James M. Calm, Emissions and environmental impacts from air-conditioning and refrigeration systems, International Journal of Refrigeration 25, pp. 293 305, 2002. [2] Samira Benhadid-Dib, and Ahmed Benzaoui, Refrigerants and their impact in the environment. Use of the solar energy as the source of energy, Energy Procedia 6, pp. 347 352, 2011. [3] Samira Benhadid-Dib, and Ahmed Benzaoui, Refrigerants and their environmental impact Substitution of hydro chlorofluorocarbon HCFC and HFC hydro fluorocarbon. Search for an adequate refrigerant, Energy Procedia 18, pp. 807 816, 2012. [4] Eric Granryd, Hydrocarbons as refrigerants - an overview, International Journal of Refrigeration 24, pp. 15-24, 2001. [5] Y.S. Lee, and C.C. Su, Experimental studies of isobutene (R600a) as the refrigerant in domestic refrigeration system, Applied Thermal Engineering 22, pp. 507 519, 2002. [6] R. Cabello, E. Torrella, and J. Navarro-Esbri, Experimental evaluation of a vapour compression plant performance using R134a, R407C and R22 as working fluids, Applied Thermal Engineering 24, pp. 1905 1917, 2004. www.ijmca.org Page 123