RECIPROCATING COMPRESSOR FOR A COOLING DEVICE

Abstract

A reciprocating compressor for a cooling device provided with a closed circuit (C) having a main branch (M), in which a first flow rate (X) of circulating coolant enters in the compressor, and a first economizer branch, or secondary branch (E), in which a second flow rate (X1) of coolant circulates under a pressure different from the pressure of the first flow rate (X) of coolant, the compressor being provided with a cylinder and a piston reciprocatingly moving in the cylinder, between a top dead center (S) and a bottom dead center (I), and having a suction duct for the entrance of the first flow rate of coolant, and a port obtained in the wall of the cylinder for the entrance of the second flow rate of coolant, in such a way that the piston exposes at least in part the first inlet port, at least during its inlet stroke, and covers the port at least during its compression stroke, wherein the first inlet port has a slit shape with the main dimension substantially transverse to the axis (A) of the cylinder.

Description

FIELD OF THE INVENTION

The present invention relates to a reciprocating compressor for a cooling device.

KNOWN PREVIOUS ART

In particular, the present reciprocating compressor is used in those cooling devices comprising a closed circuit in which a determined flow rate of coolant circulates and that is provided with a main branch and at least one economizer and/or secondary branch. In such branches of the closed circuit two defined fractions of the overall flow rate of the coolant circulate, then exiting from the compressor. Such an economizer, or secondary, branch is fluidically connected to a section of the closed circuit comprised between the condenser and the expansion valve, on the one hand, and to the cylinder of the reciprocating compressor for the re-injection, into the compressor itself, of the fraction of flow rate crossing the secondary branch, on the other hand. Still in a known way, along such a closed circuit a condenser, an expansion valve, an evaporator and the reciprocating compressor itself are fluidically connected one to the other. Yet still in a known way, the fraction of coolant circulating in the economizer, or secondary, branch, that can comprise an additional expansion valve and an heat exchanger or an additional evaporator, has a pressure value intermediate between the highest and the lowest of the circuit of the cooling device, i.e. between the pressure of the fluid to the condenser and that one to the evaporator being along the main branch.

In general, in compressors usually adopted in refrigeration devices, the exact point of the compressor in which the aforementioned fraction of flow rate coming from the secondary economizer branch is entered, can always be determined. For example, in a screw compressor, in which as it is known the pressure increases along the compressor axis according to a known law, the exact point of injection of the fraction of flow rate coming from the secondary economizer branch can always be located. The same applies also for other types of compressors such as, for example, screw or scroll compressors, although the operating principle as well as the pressure distribution inside the compression chamber are different with respect to that one of the screw compressors, however also in the scroll compressor it can always be known how great is the pressure in any point of the compression chamber.

In case of use of reciprocating compressors, i.e. provided with cylinder and piston reciprocatingly moving inside the cylinder, the pressure instead varies with time and is anytime substantially the same in the whole cylinder for every position of the piston during its inlet and compression stroke.

However, in order to allow using economizer, or secondary, branches in cooling devices employing a reciprocating compressor, in document US2014/0170003 in the name of Emerson Climate Technologies Inc. the use of cylinders is described that, beside comprising a conventional suction duct located on the header, are also provided with a side inlet port with circular section for the entrance of such a fraction of flow rate coming from the afore mentioned economizer branch at a defined intermediate pressure. At the inlet port being in the compressor cylinder a valve is located whose opening and closing is synchronized with the compressor drive shaft through a complicated mechanism consisting of at least one cam and at least one respective follower. This allows the aforementioned fraction of flow rate of coolant coming from the secondary economizer branch to be entered only shortly before a pressure slightly smaller than the pressure of the afore mentioned fraction of secondary flow rate is reached in the piston.

In order to avoid using complex synchronization systems, as those described in US2014/0170003, other solutions have been studied. In particular, in document WO-A1-2007064321 in the name of Carrier Corporation, it is taught how to implement on the compressor cylinder an inlet port with circular section that is exposed by the piston in its inlet stroke and remains covered, still by the piston, during the compression stroke of the latter. Unfortunately, with respect to reciprocating compressors having the same displacement, but free from side port, a remarkable reduction of the possible compressor work is obtained, since a part of the piston stroke is used to allow the inflow fraction of the flow rate of coolant coming from the economizer, or secondary, branch. In particular, in cases where such a fraction of flow rate is considerable, up to 50% of the overall flow rate value, the use of compressors having the same displacement as those employed in cooling devices free of economizer, or secondary, branch, becomes highly difficult. In fact, in such cases, the inlet port for the flow rate of coolant has remarkable dimensions along the axis of the cylinder with the result that a compressor having a displacement greater than those normally used, and thus with an increase of overall costs, has to be employed. Therefore, object of the present invention is to realize a reciprocating compressor that can be used in cooling devices provided with at least one economizer, or secondary, branch, but that—the performance being the same—has a displacement lower than the displacement at present employed in such cooling devices. Further object of the invention is to realize a reciprocating compressor that, in addition to achieving the afore mentioned object, is highly simple to implement, even starting from know compressors free of side port.

SUMMARY OF THE INVENTION

These and other objects are achieved by the reciprocating compressor for cooling device provided with a closed circuit having a main branch, in which a first flow rate of circulating coolant enters in said compressor, and at least one first economizer branch, or secondary branch, in which a second flow rate of coolant circulates under a pressure different from the pressure of said first flow rate of coolant, said compressor being provided with at least one cylinder and at least one piston reciprocatingly moving in said at least one cylinder, between a top dead centre and a bottom dead centre, and comprising at least one suction duct for the entrance of said first flow rate of coolant, and at least one port obtained in the wall of said cylinder for the entrance of said second flow rate of coolant, in such a way that said piston exposes at least in part said at least one first inlet port, at least during its inlet stroke, and covers said at least one port at least during its compression stroke, characterized in that said at least one first inlet port has a slit shape with the main dimension substantially transverse to the axis of said cylinder.

In practice, the presence of a first inlet port having the shape of a slit, with the main dimension, the length one, substantially transverse to the axis of the cylinder, allows a remarkable amount of flow rate of coolant coming from an economizer, or secondary, branch to enter the cylinder, without this concretely affecting the dimensions of the displacement of the compressor itself. In fact, the slit dimensions are highly limited along the axial direction of the cylinder, thus in height, whereas they are remarkably larger transversely to the cylinder axis, thus in length. As mentioned, this allows a remarkable flow rate of coolant to flow in the cylinder in the same very short time equal to the piston stroke during the opening and subsequent closing of the side port.

It has to be observed that the term slit has to be intended as any notch, of any shape, made in the cylinder wall and having a dominant dimension (also named as main dimension) with respect the other. In particular, in the present instance, the main or dominant or more relevant dimension is that one lying on a plane transverse to the axis of the compressor cylinder, thus not the slit dimension parallel to the axis of the compressor cylinder and defined as slit height.

According to the herein described embodiment, said at least one first port is arranged next to the bottom dead centre of said at least one piston and, preferably, said at least one first port has a lower side substantially flush with the bottom dead centre of said piston. Such a solution allows avoiding an excessive loss of compressor displacement and compression work, simultaneously, in its inlet and compression stroke at the side port.

According to the invention, said at least one closed circuit of said cooling device further comprises at least an additional economizer branch, or secondary branch, in which an addition flow rate of said coolant is circulating, said compressor further comprising at least one second port obtained in the wall of said cylinder for the entrance of said additional flow rate of said coolant in said at least one compressor, wherein said at least one second port has a slit shape with the main dimension substantially transverse to the axis of said cylinder and is arranged at a distance from said bottom dead centre greater than the distance at which said at least one first port is positioned, so that said piston exposes said at least one second inlet port at least during its inlet stroke, and covers said at least one port at least during its compression stroke. Such a configuration is particularly suited in case the additional flow rate coming from an additional economizer, or secondary, branch has a pressure lower than the pressure of said second flow rate coming from the economizer, or secondary, branch, and entering the cylinder of the compressor through said at least one first port.

According to the embodiment herein described, said at least one first port and said at least one second port, both having a slit shape, are substantially or mainly rectangular-shaped, i.e. the slit surface, that one facing the inner face of the compressor cylinder, has substantially the shape of a rectangle lying on the inner cylindrical surface of the compressor cylinder. Such a substantially rectangular shape, where the top or bottom side has dimensions greatly larger than those of the two height sides, i.e. along the axial direction of the compressor cylinder, could also have sides blent one to another, i.e. without sharp edges, falling however in the definition of surface having substantially a shape of rectangle lying on the inner surface of the cylinder.

Furthermore, the ratio between height and length dimensions, i.e. along the main direction, of said at least one first port and/or said at least one second port is smaller than 0.5, preferably 0.2. In fact, the Applicant tested that such dimensional values are those that allow obtaining the best performances. It has to be noted that the slit length has to be calculated along the arc of a circle of the cylinder along which the same slit extends, on a plane transverse to the cylinder axis and passing in the middle of the slit height.

In addition, the lower side of said at least one second port is flush with the upper side of said at least one first port.

According to a further embodiment, said at least one first inlet port and/or said at least one second inlet port comprises/comprise at least one functionally-combined non-return valve. Such non-return valves allow preventing the coolant entered the compressor through the first and the second port from being biased towards them in the opposite way, during the rising step of the piston, i.e. the coolant compression step.

More specifically, such at least one non-return valve is of deformable reed type and is housed in the wall of said at least one cylinder. This makes the compressor even more compact, however avoiding the presence of complicated elements employed for synchronizing the opening or closing of the side ports.

BRIEF DESCRIPTION OF THE DRAWINGS

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

FIG. 1 is a schematic view of a cooling device provided with a reciprocating compressor according to the invention;

FIG. 2 is a P-h diagram of the refrigeration cycle relating to the refrigeration device of FIG. 1;

FIGS. 3a-3d are schematic and sectional longitudinal views of the inside of the compressor cylinder during the inlet and compression steps;

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

FIG. 5 is a schematic view of another cooling device provided with a reciprocating compressor according to the invention;

FIG. 6 is a P-h diagram of the refrigeration cycle relating to the refrigeration device of FIG. 5.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS OF THE INVENTION

Referring particularly to such figures, with numeral 100 is denoted the reciprocating compressor according to the invention.

In FIG. 1 the scheme of a cooling device 200 provided with a reciprocating compressor 100 according to the invention provided with a cylinder 110 and a piston 111 reciprocatingly moving in the cylinder 110, between a top dead centre S (see FIG. 3d) and a bottom dead centre I (see FIG. 3c), is shown. In particular, the refrigeration device 200 comprises a closed circuit C in which a certain flow rate of coolant is circulating. Such a closed circuit C comprises, in its turn, a main branch M, in which a first flow rate X of coolant is circulating and enters the compressor 100 through a suction duct 109, a first secondary branch E and a second secondary branch E′. In such a first secondary branch E a second flow rate X1 of coolant circulates, whereas in the additional secondary branch E′ an additional flow rate X2 of coolant circulates. The sum of the flow rates circulating in the main branch M and in the two secondary branches E and E′ is the flow rate circulating in the closed circuit C and exiting from the reciprocating compressor 100.

According to the scheme of FIG. 1, the cooling device 200 further comprises a condenser 101, a first expansion valve 102 and a first evaporator 103. Both the first expansion valve 102 and the first evaporator 103 are located along the main branch M of the closed circuit C so that the flow rate X circulating in the same, which is given by the difference between the overall flow rate circulating in the closed circuit C and the flow rates X1 and X2 circulating in the two secondary branches E and E′, directly enters the reciprocating compressor 100 through the suction duct 109 (see FIG. 3a) being in the reciprocating compressor 100, above the cylinder 110.

The two secondary branches E and E′ each comprise a second expansion valve 130, 130′ and a respective second evaporator 140 and 140′; in practice, in each secondary branch E, E′, all the values of the circulating flow rate and the pressure and temperature will be different. In this type of configuration, in practice, the cooling device 200 is able to cool three different chambers connected to the respective evaporators 103, 140 and 140′. In particular, the second flow rate X1 and the additional flow rate X2 respectively circulating along the secondary branch E and the additional secondary branch E′ are at different temperatures and pressures. In particular, the pressure of the flow rate X1 circulating along the secondary branch E is intermediate between the pressure of the fluid at the condenser 101 and the pressure at the first evaporator 103, whereas the pressure of the additional flow rate X2 of coolant circulating in the additional secondary branch E′ has an intermediate value between the value of the fluid of the first flow rate X and that of the fluid of the second flow rate X1.

Note that in FIG. 1 the thermodynamic states of the coolant circulating in the closed circuit C of the refrigeration device 200 are denoted in brackets, with numbers from 1 to 8. Then in FIG. 2 the thermodynamic cycle made by the coolant in the device 200 is shown, with the information of the thermodynamic condition of the fluid in the corresponding points of the closed circuit C. The references 9 and 10 shown in the graph of FIG. 2 correspond to the thermodynamic state of the coolant in the compressor 100 in the inlet step (FIGS. 3b and 3c) at the opening of the second port 112 and the first port 107 that are on the wall 110a of the cylinder 110 of the reciprocating compressor 100, as it will be described later.

In FIG. 5, an additional cooling device 200′ comprising a reciprocating compressor 100 similar to that one of the embodiment shown in FIG. 1 is shown.

The cooling device 200′ comprises a closed circuit C comprising a main branch M along which a first flow rate X of the coolant is circulating at a defined pressure, a condenser 101, an evaporator 103, and a first expansion valve 102 arranged between the condenser 101 and the evaporator 103. Such a closed circuit C also comprises a first economizer branch E along which a second flow rate X1 of the coolant circulates. Such a first economizer branch E is fluidically connected to the compressor 100 and to a section 106 of the closed circuit C comprised between the condenser 101 and the expansion valve 102.

In the herein described embodiment, the closed circuit C further comprises an additional economizer branch E′ for an additional flow rate X2 of the coolant.

Still according to the herein described embodiment, the economizer branch E and the additional secondary economizer branch E′ comprise a second expansion valve 150, 150′ and at least one heat exchanger 160, 160′ with the section 106 of the closed circuit C comprised between the condenser 101 and the expansion valve 102.

According to the herein described embodiment, such a second flow rate X1 has an inlet pressure P₈ in the cylinder 110 of the compressor 100 intermediate between the pressure at the condenser P₂ and the inlet pressure in the cylinder 110, i.e. the pressure P₁ of the flow rate X of fluid entering the cylinder 110 of the compressor from the suction duct 109, during the inlet step of the compressor 100.

Note that in FIG. 5 the thermodynamic states of the coolant circulating in the closed circuit C of the refrigeration device 200′ are denoted in brackets, with numbers from 1 to 10. Then, in FIG. 6 the thermodynamic cycle made by the coolant in the closed circuit C is shown, with the information of the respective thermodynamic condition of the coolant. The references 11 and 12 shown in the graph of FIG. 6 correspond to the thermodynamic states of the coolant in the compressor 100 in the inlet step (FIGS. 3b and 3c) at the opening of the second port 112 and the first port 107 that are on the wall 110a of the cylinder 110 of the reciprocating compressor 100, as it will be described later.

According to the invention, in both the cooling devices 200 and 200′ the reciprocating compressor 100 comprises a first side port 107 obtained on the wall 110a of the cylinder 110 for the entrance of the aforementioned second flow rate X1 of coolant.

The compressor 100 further comprises a second inlet port 112 for the entrance of such an additional flow rate X2 of coolant. More specifically, the second inlet port 112 is arranged at a distance D from the bottom dead centre I of the piston 111 greater than the distance d at which the first port 107 is located. Such a distance is assessed with respect to two planes P and P1 transverse to the axis A of the cylinder 110 and passing in the middle of the height H of the port 107, 112.

According to the herein disclosed embodiment, the first inlet port 107 for the second flow rate X1 of coolant, that in the present instance is R404a, is a slit and 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. In addition, the second inlet port 112 for the entrance of such an additional flow rate X2 of coolant arranged, as mentioned afore, at a distance D from the bottom dead centre I of the piston 111 greater than the distance d at which the first port 107 is located, is also a slit. Also the second slit 112 is arranged on the wall 110a of the cylinder so that the piston exposes the second inlet port 112, during its inlet stroke, and before exposing the first inlet port 107, and covers it during its compression stroke, after covering the first port 107.

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 A 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 A of the cylinder 110 and forward direction of the milling machine orthogonal to the axis of the cylinder 110. Therefore the so obtained surface is substantially rectangular-shaped, despite the rectangle 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 110c (see in particular the dotted line shown in FIG. 4b), is 0.2. In particular, the length L has to be measured on a plane P, or P1, transverse to the axis of the cylinder A and passing in the middle of the height H of the respective slit.

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.

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 particularly, the lower side 112a of the second port 112 is flush with the upper side 107b of the first port 107.

According to the embodiment shown in the FIGS. 3a to 3d, only the second inlet port 112 comprises a functionally-combined non-return valve 180; whereas, in the embodiment shown in FIGS. 4a and 4b, both the first inlet port 107 and the second inlet port 112 comprise a functionally-combined non-return valve 180 of deformable reed type.

Such a non-return valve 180 is, in practice, dimensioned so as to deform only after a defined pressure is exceeded. In addition, such a non-return valve 180 is housed in the wall 110a of the cylinder 110 of the compressor 101 and, when in a not deformed condition, is in abutment against a pair of projections 190 and 191 contacting the outer surface 110b of the cylinder 110.

It has to be mentioned that, although a compressor 100 provided with a first port 107 and a second port 112 and, thus, a cooling device 200 or 200′ provided with a secondary, or economizer, branch E and an additional secondary, or economizer, branch E′ has been described heretofore, however a solution in which the compressor 100 is provided with at least one first port 107, but free of said at least one second port 112, and thus a cooling device 200 or 200′ provided with the only economizer, or secondary branch E, still falls within the protection scope of the present invention. In this case the first flow rate, that entering the compressor 100 through the suction duct 109, is given by the difference between the overall flow rate circulating in the closed circuit C and the only second flow rate X1.

The operation of the reciprocating compressor 100 being in the two refrigeration devices 200, 200′ respectively described in FIGS. 1 and 5, 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 113 of the compressor 100 is open to accommodate the flow rate of fluid X coming from the main circuit M, through suction duct 109 (see FIG. 3a). Subsequently, the piston 111 exposes the second port 112 from which an additional flow rate X2 of coolant coming from the additional secondary economizer branch E′ comes; due to the pressure increase, the suction valve 109 closes. The pressure of such an additional flow rate X2 of coolant is higher than the pressure being in the cylinder 110, thus resulting in a pressure increase inside the cylinder 110 (thermodynamic state 9 or 11, depending on the cooling device 200 or 200′). Of course during such a step the non-return valve 180 remains open (see FIG. 3b).

Then, the piston exposes the first port 107 thus allowing the second flow rate X1 of coolant coming from the secondary economizer branch E accessing the cylinder 110. Of course, the pressure of the second flow rate X1 of coolant coming from such an economizer, or secondary, branch E is higher than the pressure of the additional flow rate X2 of coolant and the suction pressure. Anyway, since the mixing there is an increase of the pressure in the cylinder 110 of the compressor 100 (thermodynamic state 10 or 12, depending on the cooling device 200 or 200′), 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 114 occurs. It has to be noted that during the rising of the piston 111, the non-return valve 180 placed in the part 110a of the cylinder 110 remains closed as the pressure in the cylinder 110 exceeds the pressure of the additional flow rate X2 coming from the additional, or secondary, economizer branch E′.

Lastly, note that the implementation of the first port 107 and/or the second port 112 preferably occurs through a simple milling operation, or similar technological operation, of the cylinder 110 along a plane transverse to the axis A of the cylinder 110 itself. This allows the cylinders of existing reciprocating compressors, that are free of through side port, being converted by means of a simple milling operation of the cylinder 110. In this way, such cylinders are made adapted to operate in cooling devices having at least one secondary, or economizer, branch without the need of subjecting the cylinder to complex interventions, from a technical point of view, or economically unattractive.

Claims

1. A reciprocating compressor for a cooling device provided with a closed circuit (C) having a main branch (M), in which a first flow rate (X) of coolant circulates and enters said compressor, a first economizer branch, or secondary branch (E), in which a second flow rate (X1) of fluid circulates under a pressure different from the pressure of said first flow rate (X) of coolant, and an additional economizer branch, or secondary branch (F), in which an additional flow rate (X2) of said coolant circulates, said compressor being provided with a cylinder and a piston reciprocatingly moving in said at least one cylinder, between a top dead center (S) and a bottom dead center (I), and comprising a suction duct for the entrance of said first flow rate of coolant, and a port obtained in the wall of said cylinder for the entrance of said second flow rate of coolant, so that said piston exposes at least in part said first inlet port, at least during its inlet stroke, and covers said port, at least during its compression stroke, wherein said first inlet port has a slit shape with the main dimension (L) substantially transverse to the axis (A) of said cylinder and is arranged at the bottom dead center of said piston, said compressor further comprising a second port obtained in the wall of said cylinder for the entrance of said additional flow rate (X2) of coolant in said compressor, wherein said second port has a slit shape with the main dimension (L) substantially transverse to the axis (A) of said cylinder and is arranged at a distance (D) from said bottom dead center centre greater than the distance (d) at which said first port is positioned, so that said piston exposes said second inlet port, at least during its inlet stroke, and covers said second port, at least during its compression stroke.

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

3. The compressor according to claim 1, wherein said first port and/or said at least ono second port has/have a substantially rectangular shape, lying on the inner cylindrical surface of said cylinder.

4. The compressor according to claim 3, wherein the ratio between the height (H) and length (L) dimensions of said first port and/or said second port is smaller than 0.5.

5. The compressor according to claim 1, wherein the lower side of said second port is flush with the upper side of said first port.

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

7. The compressor according to claim 6, wherein said non-return valve is of deformable reed type.

8. The compressor according to claim 7, wherein said non-return valve is housed in the wall of said cylinder.

9. The compressor according to claim 4, wherein the ratio between the height (H) and length (L) dimensions of said first port and/or said second port is smaller than 0.2.