REFRIGERATION DEVICE

Abstract

A refrigeration device having a closed circuit in which a flow rate of coolant is circulating is provided. The closed circuit has a condenser and a main branch provided with a reciprocating compressor inside which a defined flow rate of the coolant enters, from the main branch, at a defined suction pressure, of an evaporator and a first expansion valve that is arranged between the condenser and the evaporator. The closed circuit further has a first secondary economizer branch for a first fraction of flow rate of the coolant, the first secondary economizer branch fluidically connecting the compressor to a section of the closed circuit between the condenser and the first expansion valve, wherein the compressor has a first side inlet port for the entrance of the first fraction of coolant flow rate.

Claims

1. A refrigeration device having a closed circuit (C) in which a flow rate of coolant is circulating, said closed circuit comprising a condenser and a main branch (M) provided with a reciprocating compressor inside which a defined flow rate (1-X1; 1-X1-X2) of said coolant enters, from said main branch, at a defined suction pressure (P.sub.1), of an evaporator and a first expansion valve that is arranged between said condenser and said evaporator, said closed circuit further comprising a first secondary economizer branch for a first fraction of flow rate (X1) of said coolant, said first secondary economizer branch fluidically connecting said compressor to a section (106) of said closed circuit (C) comprised between said condenser and said first expansion valve, wherein said compressor comprises a first side inlet port for the entrance of said first fraction (X1) of coolant flow rate, said first fraction of flow rate having an inlet pressure (P.sub.8) so that P.sub.8−P.sub.1≦4 bar.

2. The refrigeration device according to claim 1, wherein said reciprocating compressor is provided with a cylinder and a piston reciprocatingly moving in said cylinder, between a top dead center (S) and a bottom dead center (I), said first side inlet port for the entrance of said first fraction (X1) of flow rate of said coolant being arranged at the bottom dead center of said piston, so that said piston exposes at least in part said first side inlet port, at least during its inlet stroke, and covers said first side port, at least during its compression stroke.

3. The device according to claim 1, wherein said closed circuit further comprises an additional secondary economizer branch for a second fraction of flow rate (X2) of said coolant, said compressor comprising a second inlet port for the entrance of said additional fraction (X2) of flow rate of coolant into said compressor, in which said second port is arranged at a distance from said bottom dead center greater than the distance at which said first port is arranged, said additional fraction of flow rate (X2) having an inlet pressure (P.sub.10) so that P.sub.1≦P.sub.10≦P.sub.8.

4. The refrigeration device according to claim 1, wherein said first inlet port and/or said second inlet port comprises/comprise a slit having a main dimension (L) substantially transverse to the axis (Z) of said cylinder.

5. The refrigeration device according to claim 4, wherein said slit comprises a substantially rectangular-shaped surface lying on the inner cylindrical surface of said cylinder.

6. The refrigeration device according to claim 5, wherein the ratio between the height (H) and the length (L) dimensions of said slit is less than 0.5.

7. The refrigeration device according to claim 1, wherein said first port has a lower side substantially flush with the bottom dead center of said piston.

8. The refrigeration device according to claim 7, wherein the lower side of said second port is flush with the upper side of said first port.

9. The refrigeration device according to claim 1, wherein said secondary economizer branch and/or said additional secondary economizer branch comprises/comprise a second expansion valve and a heat exchanger with said section of main branch between said condenser and said expansion valve.

10. The refrigeration device according to claim 1, wherein said secondary economizer branch and/or said additional secondary branch comprises/comprise a pipe having a cylindrical section and a fitting with said first inlet port and/or said second inlet port.

11. The device according to claim 10, wherein said cylindrical pipe is dimensioned so that to be of tuned type.

12. The device according to claim 1, wherein said first inlet port and/or said second inlet port comprises/comprise a functionally-combined non-return valve.

13. The device according to claim 12, wherein said non-return valve is of deformable reed type.

14. The device according to claim 13, wherein said non-return valve is housed in the wall of said cylinder.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

[0025] For illustration purposes only, and without limitation, several particular embodiments of the present invention will be now described referring to the accompanying figures, wherein:

[0026] FIG. 1 is a schematic view of a refrigeration device according to the invention, with two secondary economizer branches;

[0027] FIG. 2 is a P-H diagram of the refrigeration cycle used in the refrigeration device of FIG. 1;

[0028] FIGS. 3a-3d are schematic and sectional views of the inside of the compressor cylinder during the inlet and compression steps, in reference to the thermodynamic states shown in FIG. 2;

[0029] FIGS. 4a and 4b are respectively two longitudinal and transverse sectional views of the cylinder of the reciprocating compressor, with particular reference to the first and the second port obtained in the wall of the compressor cylinder;

[0030] FIG. 5a shows a schematic view of a conventional refrigeration device with reciprocating compressor and without one or more secondary economizer branches;

[0031] FIG. 5b shows a P-H diagram of the refrigeration cycle adopted in the refrigeration device of FIG. 5a.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS OF THE INVENTION

[0032] Referring in particular to such figures, the generic refrigeration device according to the invention has been denoted with numeral 100.

[0033] The refrigeration device 100 comprises a closed circuit C in which a flow rate of coolant 1 is circulating. Such a closed circuit C comprises a condenser 102 and a main branch M having a reciprocating compressor 101 provided with a cylinder 110 and a piston 111 reciprocatingly moving inside the cylinder 110, between a top dead centre S (see FIG. 3d) and a bottom dead centre I (see FIG. 3c), and inside which a defined flow rate 1-X1-X2 of the coolant enters, from said main branch M, at a defined suction pressure P.sub.1. Such a main branch M is further provided with an evaporator 103 and a first expansion valve 104 arranged between the condenser 102 and the evaporator 103. Such a closed circuit C comprises, in addition, a first secondary economizer branch 105 for a first fraction of flow rate X1 of the coolant. Such a first secondary economizer branch 105 is fluidically connected to the compressor 101 and to a section 106 of the closed circuit C comprised between the condenser 102 and the expansion valve 104. According to the invention, the reciprocating compressor 101 comprises a first side port 107 obtained on the wall 110a of the cylinder 110 for the entrance of the aforementioned first fraction X1 of flow rate of coolant.

[0034] Note that in FIG. 1 the thermodynamic states of the coolant circulating in the closed circuit C of the refrigeration device 100 are denoted in brackets, with numbers from 1 to 12. Then, in FIG. 2 the thermodynamic cycle made by the coolant in the closed circuit 100 is shown, with the information of the thermodynamic condition of the fluid at the corresponding points of the closed circuit C.

[0035] Advantageously and according to the invention, such a first fraction of flow rate X1 has an inlet pressure P.sub.8 in the cylinder 110 of the compressor 101 so that P.sub.8-P.sub.1≦4 bar, and preferably lower than 2 bar, wherein P.sub.1 is the pressure of the flow rate of the fluid 1-X1-X2 entering the cylinder 110 of the compressor 101 from the suction valve 101a, during the inlet step of the compressor 101. In practice, the Owner found that by increasing the specific volume of the fluid introduced in the cylinder through the first secondary economizer branch 105, i.e. by reducing the inlet pressure P.sub.8 to the cylinder 110 through the first side port 107 as much as possible, several advantages are achieved. Firstly, thanks to such a solution, the efficiency of the refrigeration cycle becomes greatly increased with respect to a refrigeration cycle working at the same conditions, i.e. same pressures, temperatures and same coolant. In addition, such a solution also allows greatly increasing the refrigeration load, the displacement of the employed reciprocating compressor 101 being the same. This is mainly due to the fact that, when the pressure P.sub.8 of said first fraction X1 of flow rate of coolant from the first secondary economizer branch 105 is reduced, a remarkable increase of the volumetric flow rate is obtained that, consequently, greatly increases the pressure of the cylinder 110 when enters the compressor 101 through said first port 107, thus resulting in a reduction of the compression work done by the compressor 101. Such a reduction of the work of the compressor 101 leads to a great increase of the efficiency of the whole refrigeration device 1. In addition, such a solution also allows greatly increasing the refrigeration load, the displacement of the employed reciprocating compressor 101 being the same.

[0036] According to the herein disclosed embodiment, the first inlet port 107 for the first fraction X1 of flow rate of the coolant, that in the present instance is R404a, is arranged at the bottom dead centre I of the piston 111, so that the piston exposes the first inlet port 107 during its inlet stroke and covers such a first inlet port 107 during its compression stroke.

[0037] In the herein described embodiment, the closed circuit C further comprises an additional secondary economizer branch 120 for a second fraction of flow rate X2 of the coolant. Thus the compressor 101 comprises a second inlet port 112 for the entrance of such an additional fraction X2 of flow rate of the coolant. Specifically, the second inlet port 112 is arranged at a distance from the bottom dead centre I of the piston 111 greater than the distance at which the first port 107 is located; such an additional fraction of flow rate X2 has an inlet pressure P.sub.10 so that P.sub.1≦P.sub.10≦P.sub.8, in which P.sub.10−P.sub.1≦2 bar and preferably lower than 1 bar.

[0038] Note that the aforementioned distance between the first port 107, or the second port 112, and the bottom dead centre I is measured along the axis Z of the cylinder 110 from the bottom dead centre of the piston 111 of the compressor 101 to the lower side 107a, or 112a, of the respective port.

[0039] Still according to the herein described embodiment, the first secondary economizer branch 105 and the additional secondary economizer branch 120 comprise a second expansion valve 130 and at least one heat exchanger 131 with the section 106 of the closed circuit C comprised between the condenser 102 and the expansion valve 104. At this point, for simplification purposes, a numerical example of the refrigeration device according to the invention is shown. In particular, it has to be observed that the thermodynamic cycle made by the coolant inside the closed circuit C is depicted in FIG. 2. Also in this case the numeral references located at the lines describing the thermodynamic transformations experienced by the coolant in the refrigeration device 100 are also detectable in the closed circuit C of the refrigeration device 100 shown in FIG. 1.

[0040] In the numerical example the condensation temperature is supposed to be 40° C., and the evaporation temperature −40° C. In addition, the subcooling at the outlet of the condenser is supposed to be of 2° C., whereas the overheating at the outlet of the evaporator to be of 5° C. In addition, in the herein described cycle, the overheating of the economizer vapor is supposed to be of 15° C., whereas the difference between the temperature of the subcooled fluid and the evaporation temperature to be of 5° C. Now, by using an iterative method and starting from pressure values P.sub.8 and P.sub.10 of respectively 3.0 bar and 1.55 bar of the fluid being respectively in the secondary economizer branch 105 and in the additional secondary economizer branch 120, the values of pressure (P), temperature (T), enthalpy (h), density (σ) and entropy (S) of the thermodynamic states 1, 3, 4, 5, 6, 7, 8, 9 e 10 can be determined. Subsequently, being the state 11 the thermodynamic state reached by the fluid at the mixing of vapor in the state 1 with the vapor produced in the additional economizer branch 120 at the thermodynamic state 10, it is calculated only once the fractions X1 and X2 of flow rate of the coolant in the first economizer branch 105 and in the additional secondary economizer branch 120 have been determined.

[0041] In particular, it turns out that:

X1=(h.sub.3−h.sub.4)/(h.sub.8−h.sub.4)=0.408

and

X2=(1-X1)*(h.sub.4−h.sub.5)/(h.sub.10−h.sub.5)=0.065

wherein
h.sub.3, h.sub.4, h.sub.5, h.sub.8, and h.sub.10 are the enthalpy values at the corresponding thermodynamic states visible in FIGS. 1 and 2, whereas 1 denotes the unit numerical value of the overall flow rate 1 of the coolant circulating in the closed circuit C.

[0042] Then, once the thermodynamic characteristics of the fluid at the thermodynamic state 12 have been determined, i.e. when the fluid coming from the secondary branch 105, at the thermodynamic state 8, mixes to the fluid being in the cylinder 110 at the thermodynamic state 11, the physical state 2′ relating to an isentropic compression can be calculated by fixing the value of 0.7 as the efficiency η of the compressor 101. From here, the value of the fluid at the thermodynamic state 2, i.e. exiting from the compressor 101, can be calculated.

[0043] In summary, the physical states of the fluid in the thermodynamic cycle according to the herein described embodiment, in view of the employed and afore mentioned hypotheses, are the following:

TABLE-US-00001 P T h σ S X 1 1.31 −35 347.6 6.81 1.6563 2 18.3 77.7 427.3 75.58 1.7266 3 18.3 38 256.8 978 1.1903 4 18.3 −15 179.9 1211 0.9205 5 18.3 −32 157.9 1267 0.8321 6 1.31 −40 157.9 0.8388 0.059 7 3.07 −20 256.8 1.2293 0.461 8 3.07 −5 368.3 14.58 1.6678 9 1.55 −37 179.9 0.9312 0.149 10  1.55 −22 357.5 7.62 1.6806 11  1.50 −29.8 351.2 7.63 1.6580 12  2.74 −6.6 367.7 12.99 1.6744  2′ 18.3 62.4 409.4 83.35 1.6744

[0044] In view of such values the coefficient of performance, or more commonly known with the acronym COP, is the following:

COP=[(1-X1-X2)*(h.sub.1−h.sub.6)]/[h.sub.2−(1−X1−X2)*h.sub.1−X1*h.sub.8−X2*h.sub.10]=1.42

wherein
h.sub.1, h.sub.2, h.sub.6, h.sub.8 and h.sub.10 are the enthalpy values of the corresponding thermodynamic states that can be seen in FIGS. 1 and 2.

[0045] On the contrary, in case of conventional refrigeration device 300 shown in FIG. 5a, i.e. provided with the condenser 102′, expansion valve 104′, evaporator 103′ and reciprocating compressor 101′ and free of secondary economizer branches, and whose thermodynamic cycle is depicted in FIG. 5b, and starting from the same working hypotheses, i.e. same condensation temperature, outlet temperature at the condenser, evaporation temperature, overheating at the evaporator outlet, entropic efficiency of the compressor, and coolant, the following values in the various thermodynamic states shown in FIGS. 5a and 5b would be obtained:

TABLE-US-00002 P T h σ S 1 1.31 −35 347.6 6.81 1.6536 2 18.3 56.7 402.5 87.01 1.6536 3 18.3 76.5 426.0 76.06 1.7229 4 18.3 38 256.8 978 1.1703  2′ 1.31 −40 256.8 12.40

[0046] Hereupon, the following coefficient of performance would be obtained:

COP′=(h.sub.4)/(h.sub.2−h.sub.1)=1.16

[0047] In practice, thanks to the herein described solution, a COP is obtained that is 22.4% greater than the COP′ that could be obtained by a conventional refrigeration device 300 however operating at the same thermodynamic conditions of that one according the invention. In practice, the energy efficiency of the refrigeration device 100 according to the invention is greatly improved.

[0048] In addition, by making further considerations on the refrigeration load of the compressor in the two afore compared refrigeration devices, i.e. the refrigeration device 100 and the refrigeration device 300, and in the light of the displacement between the two reciprocating compressors 101 and 101′ being substantially similar, this hypothesis being close to the truth, the following results will be obtained:

Q/Q′=[σ.sub.12(1-X1-X2)*(h.sub.1−h.sub.6)]/[σ.sub.1′(h.sub.1′−h.sub.4′)]=2.1

Wherein:

[0049] Q is the refrigeration load of the refrigeration device 100 according to the invention;
Q′ is the refrigeration load of the refrigeration device 300 according to the scheme of FIG. 5a;
σ.sub.12 is the fluid density in the refrigeration device 100 and in the thermodynamic state 12;
σ.sub.1′ is the fluid density in the refrigeration device 300 and in the thermodynamic state 1;
h.sub.1′ is the fluid enthalpy in the refrigeration device 300 and in the thermodynamic state 1;
h.sub.4′ is the fluid enthalpy in the refrigeration device 300 and in the thermodynamic state 4.

[0050] In practice, the refrigeration load of a compressor 101 operating in a refrigeration device 100, in which the pressure of the first fraction of flow rate P.sub.8 entering the compressor 100 is such that P.sub.8−P.sub.1≦4 bar and in which the pressure of the second fraction of flow rate P.sub.10 entering the compressor 100 is such that P.sub.10−P.sub.1≦1 bar, is twice than that one of a reciprocating compressor 101′ that operates in a refrigeration device 300 of known art and has the same displacement.

[0051] It has to be noted that the herein described embodiment 100 comprises a first economizer branch 105 and a second economizer branch 120, however an embodiment free of the additional economizer branch 120 still allows reaching the objects of the present invention and is, therefore, included in the protection scope of the present invention. In this case, the flow rate entering the compressor 100 would be given by the difference between the total flow rate 1 and that one of the fraction of flow rate X1 to the economizer branch 105, and would be denoted by the reference 1−X1 rather than 1−X1−X2, as done heretofore.

[0052] In particular, according to the herein described embodiment, both the first inlet port 107 and the second inlet port 112 comprise a slit whose main dimension L is arranged on a plane P, P1 substantially transverse to the axis Z of the cylinder 120.

[0053] In particular, both the first inlet port 107 and the second inlet port 112 comprise a slit whose main dimension L is substantially transverse to the axis Z of the cylinder 110. In particular, the slit has a substantially rectangular-shaped surface, lying on the inner surface 110c of the cylinder 110, thus along an arc of a circle of the cylinder 110. More specifically, for example such a surface is obtained through a cutting by milling machine of the wall 110a of the cylinder 110, obtained with the rotation axis of the milling machine parallel to the axis Z of the cylinder 110 and forward direction of the milling machine orthogonal to the axis Z of the cylinder 110, in radial direction. Therefore the so obtained surface is substantially rectangular-shaped, despite the sides are not reciprocally connected by sharp edge, but are blent one to the other. Preferably, the ratio between the H height dimension and L length dimension (also main dimension), the latter being measured along the arc of a circle traveled by the slit along the inner surface of the cylinder 110b (see in particular the dotted line shown in FIG. 4b), is 0.2. In particular, the length has to be measured on a plane P, or P1, transverse to the axis of the cylinder Z and passing in the middle of the height H of the respective slit.

[0054] Note that, anyway, any slit having a dimensional ratio of height H to length L smaller than 0.5 still falls within the protection scope of the present invention. In addition it has to be noted that the slit, i.e. the surface extending on the inner face 110c of the cylinder 110, has lower and upper sides blent to the respective connecting sides, since it follows the shape of the wall 110a of the cylinder 110 itself.

[0055] In particular, as visible in FIGS. 3a to 3d, the first port 107 has a lower side 107a substantially flush with the bottom dead centre I of the piston 111. More specifically, the lower side 112a of the second port 112 is flush with the upper side 107b of the first port 107.

[0056] According to the herein shown embodiment, both the first secondary economizer branch 105 and the additional secondary economizer branch 120 have a pipe 132 with a cylindrical section and a fitting 133 converging to the respective inlet port, i.e. to the first port 107 and to the second port 112. In particular, such a cylindrical pipe 132 is dimensioned so that to be of tuned type. It has to be noted that a similar convergent fitting (not shown herein) is also placed between the pipe 132 and the outlet of the heat exchanger 131 located downstream of the same pipe 132.

[0057] According to the embodiment shown in the FIGS. 3a to 3d, only the second inlet port 112 comprises a functionally-combined non-return valve 140; on the contrary, in the embodiment shown in FIGS. 4a and 4b, both the first inlet port 107 and the second inlet port 112 have a functionally-combined non-return valve of deformable reed type.

[0058] Such a non-return valve 140 is in practice dimensioned so as to deform only after a defined pressure is exceeded. Furthermore, such a non-return valve 140 is housed in the wall 110a of the cylinder 110 of the compressor 101.

[0059] The operation of the reciprocating compressor being in the refrigeration device 100 is explained in FIGS. 3a to 3d. In practice, during the inlet step of the compressor, i.e. when the piston 111 of the compressor 101 slides downwards from the top dead centre S to the bottom dead centre I, the suction valve 101a of the compressor is open to accommodate the flow rate of fluid 1-X1-X2 coming from the main circuit M and in the thermodynamic state 1 (see FIG. 3a). Subsequently, the piston 111 exposes the second port 112 from which a second fraction X2 of flow rate from the additional secondary economizer branch 120 comes; due to the pressure increase, the valve 101a closes. The pressure P.sub.10 of such a second fraction X2 of flow rate is higher than the pressure P.sub.1 being in the cylinder 110, thus resulting in a pressure increase inside the cylinder 110 (thermodynamic state 11). Of course during such a step the non-return valve 140 remains open (see FIG. 3b).

[0060] Then, the piston exposes the first port 107 thus allowing the access of the first fraction X1 coming from the secondary economizer branch 105 to the cylinder 110. Of course, the pressure P.sub.8 of the first fraction X1 of flow rate coming from such a first economizer branch 105 is higher than the pressure of the second fraction X2 of flow rate and than the suction pressure P.sub.1, however, advantageously, such a pressure P.sub.8 does not exceed the pressure of the flow rate 1-X1-X2 entering the compressor 101 and coming from the main branch M for more than 4 bar. In any case, since the mixing there is an increase of the pressure in the compressor 101 (thermodynamic state 12), before the latter starts its compression stroke. Subsequently, the piston 111 rises again and compresses the fluid in the cylinder 110, until reaching the top dead centre S. When the pressure in the cylinder exceeds the condensation pressure, the opening of the exhaust valve 101b occurs. It has to be noted that during the rising of the piston 111, the non-return valve 140 placed in the part 110a of the cylinder 110 remains closed as the pressure in the cylinder exceeds the pressure of the flow rate coming from the additional secondary economizer branch 120.