RECIPROCATING COMPRESSOR WITH HIGH FREEZING EFFECT

Abstract

A reciprocating compressor (1) for a coolant comprising at least one drive shaft (3) actuated by a motor (2); one compression piston (9) for the coolant (R) slidable in a respective compression chamber (5) and actuated by said shaft (3); suction system (13) to supply the coolant (R) to be compressed in said compression chamber (5) at a suction pressure (P1); a delivery system (15) to inject the compressed coolant (R) in an external refrigerating circuit or machine. The compressor presents a distribution system (19, 21) to inject in the compression chamber, in phase with the movement of the piston, coolant at a pressure intermediate between the suction pressure and the delivery pressure.

Description

CROSS REFERENCE TO RELATED APPLICATIONS

This application is a United States National Phase application of International Application PCT/IT2010/000129, the entire contents of which are incorporated herein by reference.

FIELD OF THE INVENTION

The present invention relates to a reciprocating compressor, in particular, although not exclusively, of the hermetic or semi-hermetic type for cooling plants, for example refrigeration plants.

BACKGROUND OF THE INVENTION

The use of hermetic or semi-hermetic compressors is generally known to feed a refrigerating circuit of cold store facilities comprising a plurality of refrigeration units (such as for example displayers, cold-storage rooms, refrigerated cabinets) for different types of products that must be stored according to different climatic conditions. These facilities can differ also substantially in the required refrigeration capacity and in engineering complexity according to the particular use requirements.

Hermetic or semi-hermetic reciprocating compressors comprise, in their main characteristics, a case, inside which an electric engine is housed to move a drive shaft connected to a plurality of pistons sliding in compression cylinders suitable to compress a coolant to perform the refrigerating cycle.

Compressors for compression cycles have been developed to improve the yield of the refrigerating cycle, which provide for sucking a first quantity of coolant in the suction chamber and then injecting in the cylinders an additional quantity of coolant at one or more pressure levels intermediate between the suction pressure and the delivery pressure. The refrigerating cycles operating according to this general principle are called “Voorhees cycles” or “steam multiple-compression refrigerating cycles”.

FIG. 1 shows a traditional pressure-volume curve, where on the axis of abscissas the position of a compression piston head or crown is indicated, measured from the bottom dead center (BDC), and on the axis of ordinates the pressure inside the cylinder is indicated. In particular, a typical pressure-volume cycle is shown there, where the quadrilateral ABCDA represents the typical trend of a traditional reciprocating compressor, where the segment AB corresponds to the piston expansion phase, the segment BC corresponds to the suction phase, the segment CD corresponds to the compression phase and the segment DA corresponds to the discharge phase. In FIG. 1 the quadrilateral ABGEA represents also the typical trend of a “Voorhees cycle”, wherein the segment AB corresponds to the expansion phase, the segment BG to the suction phase, the segment GE to the compression phase, and the segment EA to the discharge phase. The ABGEA cycle differs from the ABCDA cycle mainly at the end of the suction phase and in the compression phase CF, as the pressure is increased, as well as the quantity of fluid in the cylinder thanks to the injection of additional fluid.

Theoretically, a compressor operating according to a multiple compression cycle therefore increases, with the same construction characteristics and boundary conditions, the refrigerating capacity and the energy efficiency relative to a compressor operating according to a traditional cycle.

U.S. Pat. No. 1,821,248, GB 28031 A A.D. 1910, GB190504448, and GB793864A, all by Voorhees, describe some types of compressors using this “Voorhees Cycle”. In particular, U.S. Pat. No. 1,821,248 and GB 28031 A A.D. 1910 describe systems for modifying a single-stage compressor and adapting it for operation on refrigerating plants with more evaporation pressure levels. GB190504448 and GB793864A describe multiple effect compressors, which provide for an injection system for injection through one or more auxiliary ports controlled by the movement of the main piston and actuated by a rotary valve moved by a gear drive system that takes the motion from the main shaft.

These multiple effect compressors have been widely used in the first decades of the past century in big ammonia refrigeration plants; those machines operated with very low rotation speeds, about 100 rpm or even less; the quantity of fluid at intermediate pressure that was possible to inject at these speeds was high, and therefore refrigerating capacity and energy efficiency increase was significant. On the contrary, the modern refrigerating compressors operate at high speeds, typically at 1450 rpm and above, and this entails a high translation speed of the piston inside the cylinder; applying the solutions indicated by Voorhees in the previously mentioned patents on compressors operating at high rotation speeds leads to an extreme reduction of the quantity of fluid at intermediate pressure which can be injected in the cylinder.

As a result, the segment GE in FIG. 1, representing the compression phase, is very similar to the segment CD related to a traditional compressor, and this significantly reduces the advantages deriving from the refrigerating capacity and energy efficiency increase. In addition to this, these compressors are very complex from a mechanical point of view, and, with the high rotation speeds of the modern compressors, they are somewhat unreliable, difficult to be used and maintained.

Substantially, compressors for refrigerating plants operating according to the multiple effect refrigeration cycles, or Voorhees cycles, currently are not widely used, due to the above mentioned difficulties.

Therefore, despite the developments of technology the problem currently exists and there is the need for reciprocating compressors, which are more versatile and efficient to be used than the current ones, and which are at the same time sufficiently reliable.

SUMMARY OF THE INVENTION

According to an aspect, the object of the present invention is to realize some improvements to a reciprocating compressor so that it is more efficient, simple and economical to be constructed and used than the current compressors, overcoming, completely or partially, one or more of the above mentioned disadvantages.

These objects and advantages are substantially obtained with a compressor as claimed in claim 1. Characteristics and particularly advantageous embodiments of the present invention are indicated in the dependant claims.

Practically, the invention provides for a reciprocating compressor for a coolant, comprising at least one piston slidable in a compression chamber, actuated by a drive shaft, wherein a supply ports exits in the compression chamber for feeding a coolant at a pressure intermediate between the delivery pressure and the suction pressure of the compressor, and wherein to the supply port a distribution system is associated, to control the supply of the coolant in synchronous manner with the position of said piston.

In some embodiments, the distribution system comprises a mechanical distributor, which opens and closes an aperture for supplying the coolant at intermediate pressure towards the compression chamber. In some embodiments the distributor is slidable in a distribution chamber in a synchronous manner with the piston or pistons of the compressor. The distributor can be associated to an arrangement of ducts corresponding, in number and position, to the number and to the reciprocal phase of the pistons of the compressor, so as to control opening and closing of the respective supply ports in the compression chambers.

With a distributor slidable in a distribution chamber it is possible to obtain opening and closing of the supply port for the coolant at the intermediate pressure in a manner synchronized with the position of the piston, and therefore the fluid at the intermediate pressure is injected in the compression chamber in a phase of the compression cycle wherein the fluid under compression has a pressure compatible with that of the fluid coming from the supply port. The system is efficient and reliable also at high rotation speeds that are typical of the modern compressors of the refrigerating plants.

The distributor can be actuated through a connecting system for cinematic connection to the drive shaft, to obtain, simply and reliably, timing between the distributor and the position of the piston or pistons. In some embodiments the cinematic connection can comprise a cam associated with the drive shaft and a tappet associated with the distributor. The tappet can be provided with a spring, which maintains it into contact with the cam.

In other preferred embodiments of the invention, the cinematic connection between drive shaft and distributor can be obtained with a rod-crank mechanism.

In possible embodiments, the distributor presents at least one passage duct to put the distribution chamber into fluid connection with the compression chamber or chambers the respective supply port or ports.

Advantageously, the distributor is suitable to open gradually the supply port alternatively in both the directions of its stroke when the passage duct crosses the supply port or passes near it; in this way it is possible to put the distribution chamber into fluid communication with the compression chamber. The distributor furthermore closes the supply port alternatively in both the directions of its stroke thanks to its outer surface or shell.

In an advantageous embodiment of the present invention, the passage duct presents at least one entrance aperture on the head of the distributor and at least one exit aperture arranged at an intermediate height on the shell.

Further embodiments are also possible according to particular use requirements, for example it is possible to obtain one or more exit apertures to put more fluid in a same compression chamber or to inject fluid in different compression chambers, or other else.

In an advantageous embodiment of the invention, the exit aperture of the passage duct is arranged in a compartment obtained on the side shell of the distributor, and in this compartment a shaped seal sliding block can be inserted, slidable in radial direction and presenting a through channel in correspondence of the exit aperture. This seal sliding block is designed for being pushed against the wall of the distribution chamber by the fluid distribution pressure, so as to increase at least partially the seal, decreasing the clearance and the leakage of the fluid secondary flows between the distributor and the same chamber.

In a particular embodiment of the invention, a backflow valve can be provided in the passage duct of the distributor, in this way the mechanical complexity and the cost of the system is increased, but the probability of a backflow of the fluid at high pressure decreases.

When the distributor is actuated by the drive shaft controlling the piston or pistons, preferably the distributor and the piston are mutually out phased so as to open gradually the supply port when the piston is near the bottom dead center or is in the compression phase, to avoid fluid at the distribution pressure being fed during the piston suction phase, but it is also possible to feed the fluid at the distribution pressure also during the suction phase, even if with a decreased energy efficiency.

In a possible embodiment of the invention, the compressor is of the two- cylinder type with a pair of compression pistons slidable in respective compression chambers, each presenting at least one supply port for the fluid connection with at least one distribution chamber, where the distributor is arranged. This latter is preferably arranged between the two compression chambers.

In this case the distributor passage duct is shaped so as to feed the fluid in the two compression chambers according to the timing of the two pistons and it can be obtained according to different conformations based upon particular use or construction requirements of the compressor. This duct can be obtained for example with a single entrance aperture on the head of the distributor and a plurality of exit apertures corresponding to each supply port, or more exit apertures can be provided for each supply port.

These two pistons can be mutually out-of-phase and the distributor can be out-of-phase relative to them by an intermediate angle equal to nearly the half of the timing thereof.

Clearly, a different number of pistons and/or distributors can be provided, arranged in various manner (in-line, V-shaped, or other else) in a compressor according to the requirements of the plant to be supplied.

An advantage of some embodiments of the compressor according to the present invention is the fact that it has high energy efficiency and refrigerating capacity, as the compressor operates a greater quantity of fluid with less leakages.

A further advantage is that it is possible to reduce the end compression temperature of the fluid, as the temperature of the injected fluid at distribution pressure is lower than the temperature of the fluid in the compression chamber; in some use conditions it is therefore possible to use a one-stage compressor according to the present invention instead of a traditional two-stage compressor, thus obtaining a considerable saving both in construction and maintenance

The various features of novelty which characterize the invention are pointed out with particularity in the claims annexed to and forming a part of this disclosure. For a better understanding of the invention, its operating advantages and specific objects attained by its uses, reference is made to the accompanying drawings and descriptive matter in which preferred embodiments of the invention are illustrated.

BRIEF DESCRIPTION OF THE DRAWINGS

In the drawings:

FIG. 1 is a pressure- volume diagram, on which a traditional refrigerating cycle, a multiple effect refrigeration cycle or Voorhees cycle, and a cycle according to the present invention are indicated;

FIG. 2 is a view in vertical section of a reciprocating compressor according to an embodiment of the present invention;

FIG. 3 is an enlarged sectional view of the distributor of the compressor of FIG. 2;

FIG. 4 is a side view of the distributor of FIG. 3;

FIG. 5 is a side view of a piston of the compressor of FIG. 2;

FIG. 6 is an enlarged view of the head of the compressor of FIG. 2 in one position during the operation cycle;

FIG. 7 is an enlarged view of the head of the compressor of FIG. 2 in another position during the operation cycle

FIGS. 8 to 10 are diagrams of the timing of some components of the compressor of FIG. 2 according to some embodiments of the present invention;

FIG. 11 is a view of a refrigerating circuit, in which the compressor of FIG. 2 is used;

FIG. 12 is a semi-logarithmic enthalpy-pressure diagram referred to the refrigerating circuit of FIG. 11;

FIG. 13 is a view of a further refrigerating circuit, in which the compressor of FIG. 2 is used; and

FIG. 14 is a semi-logarithmic enthalpy-pressure diagram referred to the refrigerating circuit of FIG. 13.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

In the drawing, a compressor according to an embodiment of the present invention is of the semi-hermetic type and is indicated with number 1 (see FIG. 2) and comprises a casing or case 1A, in which an engine compartment 1B is obtained, inside which an electric motor 2 is arranged and connected mechanically to a crank drive shaft 3.

A second compartment 1C, adjacent to the engine compartment 1B and divided from it by means of a dividing wall 1D, is shaped at the bottom so as to produce an oil pan for the lubricant L, in fluid connection with the lower part of the engine compartment 1B through a passage hole 1E obtained on the diving wall 1D. The lubricant L is centrifuged by the disk 1F integral with the drive shaft 3, passes from the pan of the compartment 1C to the tray 1G, and through the hole 1H obtained in the drive shaft 3 lubricates the bronze bushes 1M and 1N and the connecting rods 9A, 11A, and 19A.

In the second compartment 1C two compression chambers 5 and 7 are obtained, inside which two compression pistons 9 and 11 respectively are housed in a slidable manner, connected to the drive shaft 3 by means of a connecting rod 9A and 11A respectively. The pistons 9, 11 are suitable for compressing a coolant R in a refrigerating machine or system, in which the compressor is inserted.

The compression chambers 5 and 7 end with an upper head 6, to which are associated a suction system 13, to supply the coolant R to each of them, and a delivery system 15, to send the compressed fluid R in the outer refrigerating machine.

In this embodiment, the suction system 13 is provided with inlet compartments 13A and 13B obtained in the head 6 and presenting respective inlet valves 13C and 13D to make the fluid selectively enter in the compression chambers 5 and 7 respectively, see also the description with reference to FIGS. 6 and 7.

The coolant R at a suction pressure P is supplied to the compressor 1 from a side entrance 1L in the engine compartment 1B, passes in an entrance 13E of a inlet duct and exits from exits 13F and 13G of such duct into the inlet compartment 13A and 13B respectively.

The delivery system 15 is provided with delivery compartments 15A and 15B obtained in the head 6 and presenting respective delivery valves 15C and 15D to open selectively the exit for the fluid from the compression chamber 5 and 7 respectively.

It is clearly apparent that this compressor 1 is described just by way of example, as it may be of any other type suitable to the purpose, for instance it can present a different number of compression chambers arranged in various manner, inline or V-shaped, with one or more drive shafts.

According to the present invention a distributor 19 is provided, slidable in a distribution chamber 21 and presenting at least one passage duct 23 (see FIG. 3) inside itself, to put into fluid connection the upper part of the distribution chamber 21 with each compression chamber 5 and 7 through a side supply port 25 and 27 respectively. In the embodiment shown in FIG. 2 the distribution chamber 21 is arranged between the two compression chambers 5 and 7; it is however clearly apparent that this distribution chamber 21 can be obtained and arranged in different manners according to particular construction or use requirements.

An inlet system 14 supplies the coolant R in the distribution chamber 21 at a distribution pressure P2 greater than the suction pressure P1, as described in greater detail hereunder.

FIG. 3 shows an enlarged section of the distributor 19, where it should be noted in particular that it presents a substantially cylindrical shape; it is clearly apparent that the shape and the dimensions of the distributor 19 represented herein are not limiting, as it can be produced in any manner useful to the purpose.

FIG. 3 furthermore shows the passage duct 23, which presents an entrance aperture 23 A on the head 19T and two exit apertures 23B and 23C at different heights on the side shell 19M of the distributor 19, so as to put into fluid communication the upper part of the distribution chamber 21 with the supply port 25 and respectively 27 (and therefore with the compression chambers 5 and 7) according to the position of the distributor 19, as described hereunder.

In the shown embodiment, the passage duct 23 is obtained with a channel 23D inside, and nearly coaxial with, the distributor 19, from which the exit apertures 23B and 23C extend radially.

In this way the distributor is particularly simple and inexpensive to be produced; it is also possible that the passage duct 23 is obtained in a different manner according to particular construction or use requirements, such as for instance with two or more inner channels and/or two or more entrance and/or exit apertures. Furthermore, it is also possible to provide for two or more supply ports 25, 27 arranged horizontally, vertically or in other manner for a compression chamber or for more compression chambers according to particular requirements.

In a particularly advantageous embodiment, the exit apertures 23B and 23C exit in respective compartments 19A and 19B obtained on the side shell 19M of the distributor 19, in which can be inserted shaped seal sliding blocks 29 slidably in radial direction, presenting a through channel 29A close to the exit aperture 23B and 23C. These sliding blocks 29 are pushed by the distribution pressure P2 of the coolant R against the wall of the distribution chamber 21, so as to decrease the clearance, the leakage and the secondary flows of the coolant R and to increase the seal between the distributor 19 and the chamber 21.

Advantageously, the through channel 29A of each seal sliding block 29 can present a diameter slightly smaller than that of the corresponding exit aperture 23B and 23C, so as to hinder the backflow of the coolant R and to increase the thrust on the seal sliding block.

In FIG. 3 it should be furthermore noted that the distributor 19, in this advantageous embodiment, provides for at least one upper seat 19C for a dry seal obtained between the head 19T and the seal sliding blocks 29, and at least one lower seat 19D for another seal obtained below the seal sliding blocks 29.

In this embodiment, the distributor 19 is actuated by the drive shaft 3 and is connected to it by means of a connecting rod 19L (see FIG. 2) using a plug, inserted in a through hole 19S below the lower seat 19D; however, it is clear that the distributor 19 can be actuated by any other suitable mechanism, provided that in synchronous manner with the motion of the pistons 9,11.

FIG. 4 shows a side view of the distributor 19 of FIG. 3, where it should be in particular noted the upper and lower seats 19C and 19D, the through hole 19S and one of the sliding blocks 29 of substantially rectangular shape and slightly concave, to follow the perimeter of the shell 19M.

FIG. 5 shows the compression piston 9 (completely similar to the piston 11), which presents a head 9T and three upper seats 9S near its head 9T for respective scrapers, a through hole 9F for a plug of the connecting rod 9A and, below the hole 9F, a lower seat 9B for a lower seal.

FIGS. 6 and 7 show a sectional enlargement of the head 6 of the compressor 1 of FIG. 2 in two different positions during the functioning cycle, wherein it should be noted in particular the suction system 13, comprising the inlet compartments 13A and 13B with the inlet valves 13C and respectively 13D to make the fluid selectively enter in the compression chamber 5 and respectively 7, and the delivery system 15, comprising a delivery compartment 15A and 15B with the delivery valves 15C and respectively 15D to open selectively the exit for the fluid from the compression chamber 5 and respectively 7.

Advantageously, the inlet system 14, for supplying the coolant in the distribution chamber 21 at a distribution pressure P2 greater than the suction pressure PI but lower than the delivery pressure P3, comprises a distribution compartment 14A obtained in the head 6, this compartment presenting a first entrance aperture 14B for the fluid connection with the distribution chamber 21 and a second entrance aperture 14C for the fluid connection with an external high pressure supply circuit, see the description below with reference to FIG. 11.

It should be noted that the first and second aperture 14B and 14C of the distribution compartment 14 do not present valves, but they are always open.

In FIGS. 6 and 7 it should be furthermore noted that the supply ports 25 and 27 present entrances 25A and 27A arranged in intermediate positions along the stroke of the distributor 19 in the distribution chamber 21. The exits 25B and 27B are arranged in intermediate positions between the positions taken by the head or crown of the pistons 9 and 11 corresponding to the positions of bottom dead center (BDC) and of top dead center (TDC).

In the illustrated embodiment, the compression chambers 5 and 7 and the distribution chamber 21 are arranged adjacent to each other substantially at the same height and present a similar shape, and the supply ports 25, 27 open (entrances and exits 25A, 27A, and 25B, 27B) at an intermediate height in the respective compression chambers of the distributor.

It is clear that the entrances 25A, 27A and the exits 25B, 27B of the supply ports 25, 27 can be in different number, shape, and dimensions, and they can be furthermore arranged at different heights in the chambers 5, 7, and 21 according to particular construction or use requirements; the position and the number of the compression and supply chambers 5, 7, and 21 can vary too, according to particular requirements, as explained above. The supply ports 25, 27 can be obtained, for instance, at different heights from each other, whilst the exit apertures 23B and 23C of the passage duct 23 can be obtained at the same height.

FIG. 6 shows a configuration, wherein the piston 9 is substantially at the bottom dead center and is going to move upwards (arrow F1); the exit 25B of the supply port 25 is open, whilst the entrance 25A is going to be put into communication with the distribution compartment 14A through the duct 23 of the distributor 19 moving downwards (arrow F2).

Instead, the piston 11 is substantially at the top dead center, the exit 27B of the supply port 27 is closed by the shell of the piston 11, whilst the entrance 27A is closed by the shell of the distributor 19 which is going to move downwards (arrow F3).

FIG. 7 shows the position of the pistons and of the distributor after a 180° turn of the drive shaft 3 relative to that of FIG. 6, wherein the piston 9 is substantially at the top dead center and is going to move downwards (arrow 4), the exit 25B of the supply port 25 is closed by the shell of the piston 9, whilst the entrance 25 A is closed by the shell of the distributor 19.

Instead, the piston 11 is substantially at the bottom dead center and it is going to move upwards (arrow F6), the exit 27B of the supply port 27 is open, whilst the entrance 27A is going to be completely opened by the distributor 19 moving upwards (arrow F5).

FIG. 8 shows a diagram, in which on the axis of abscissas the crank angle is indicated and on the axis of ordinates the distance of the piston 9 (curve B9), of the piston 11 (curve B11) and of the distributor 21 (curve B21) from the respective bottom dead center is indicated. In this diagram it should be noted in particular that the pistons 9 and 11 are advantageously mutually displaced by 180°, whilst the distributor 21 is displaced by 90° relative to each of them.

It is clear that this displacement can be varied according to the number of pistons and/or distributors or according to particular construction and use requirements of the compressor and of the refrigerating plant.

FIG. 9 shows a diagram, in which on the axis of abscissas the crank angle is indicated and on the axis of ordinates the passage opening or area (in square millimeter, mm) of the entrance 25A and of the exit 25B of the supply port 25 as a function of the position of the aperture 23B on the distributor 21 and respectively of the head or crown of the piston 9. The trend of the net flow sections or passage areas are represented by the curves U25A and U25B respectively for the entrance and the exit 25A and 25B. A similar diagram can be drawn with reference to the compression chamber 7 as a function of the given phasing angle.

In the case of the graph of FIG. 9, the entrance 25A is obtained in an intermediate position relative to the stroke of the distributor 19 in the chamber 21, whilst the exit 25B is arranged so as to be completely uncovered when the piston 9 is in the bottom dead center.

It should be noted in particular that the exit 25B starts to open (point N1 of the curve U25B with a nearly 130° crank angle) when the piston 9 is sliding towards its BDC, and it opens completely when the piston 9 has achieved the BDC (point N2 at about 180°). By furthermore increasing the crank angle, the piston 9 starts its stroke upwards towards the TDC, gradually closing the exit 25B, which remains closed from the point N3 at about 230° until the subsequent downstroke of the piston.

Instead, the entrance 25A starts gradually to open (point N4 of the curve U25A at about 180°) close to the exit 23B of the distributor 21, which slides towards its BDC, it completely opens (point N5 at about 210°) and then gradually closes (point N6 at about 270°).

The injection of fluid from the distribution chamber 21 to the compression chamber 5 therefore occurs between the crank angle of about 180° and the crank angle of about 230°, i.e. when both the passage opening of the entrance 25A and that of the exit 25B of the port 25 are at least partially open. The net passage opening is maximum at the point N7. It should be noted that the entrance section 25A starts again to open gradually (point N6 at about 270°), it opens completely again (point N8 at about 330° and closes gradually (point N9 at about 360°). However, during this second opening phase the exit opening 25B is null, as it is completely closed by the shell of the piston 9. The seal segment mounted in the lower seat 9B (see FIG. 5) limits the entity of fluid leakages between the shell of the piston 9 and the walls of the compression chamber 5.

FIG. 10 shows a further diagram, in which the axis of abscissas and the axis of ordinates show the same variables as the diagram of FIG. 9. The difference from the diagram of FIG. 9 is that the exit 25A is obtained in the chamber 5 in an intermediate position relative to the stroke of the piston 9, and therefore the exit 25B remains open longer. This substantially corresponds to the configuration of the compressor of FIGS. 2, 6, and 7. Furthermore, on the diagram of FIG. 10 the opening and closing curves are indicated of the entrances and the exits 25A, 25B and 27A, 27B, i.e. for both the pistons 9 and 11. The curves are indicated with the letter U, followed by the reference of the opening (ex. 25A; 25B) to which they refer. The curves for the openings 25A, 25B are represented with a continuous line, whilst the curves for the openings 27A, 27B are indicated with a dashed line.

In particular (and with initial reference to the compression chamber 5), according to this configuration the piston 9 starts to open the exit 25B (point M1 of the curve U25B with a crank angle of about 120°) during its stroke downwards from the TDC. Subsequently, the piston 9 continues to go down, completely opening the exit 25B (point M2I at about 130°). By further increasing the crank angle, the piston 9 continues its descending stroke towards the BDC and the exit 25B remains open. The piston 9 arrives at the bottom dead center, and then it inverts its stroke until it meets again the exit 25B, closing it gradually (point M2II at about 220° until M3 at about 240°).

Instead, the entrance 25A starts to open (point M4 of the curve U25A at about 180°) close to the exit 23B of the distributor 19, which slides towards the bottom dead center until the maximum opening (point M5 at about 220°) and then gradually closes (point M6 at about 240°).

In this case, the fluid injection from the distribution chamber 21 to the compression chamber 5 therefore occurs between the crank angle of about 180° and the crank angle of about 240°, i.e. when both the passage opening of the entrance 25A and that of the exit 25B of the port 25 are at least partially open. The passage section of fluid injection is maximum in the point M5, where at the same time the passage opening of the entrance 25A and of the exit 25B are maximal.

It should be furthermore noted that in this case the exit 25B remains completely open much more longer (from the point M2I to M2II), and the quantity of injected fluid is therefore greater than in the configuration described in FIG. 9.

In FIG. 10 with dashed lines are furthermore indicated the curves U27A and U27B, representing the passage opening or area of the entrance 27A and of the exit 27B of the supply port 27 as a function of the stroke of the distributor 19 and respectively of the piston 11, not described in detail for the sake of simplicity.

FIG. 1, together with the cycle of a traditional compressor and that of a “Voorhees cycle” compressor, described above with reference to the prior art, also shows the diagram of the compressor in the configuration shown in FIGS. 2, 6, and 7 and represented by the quadrilateral ABCFA, wherein the segment AB corresponds to the piston expansion phase, the segment BC to the suction phase, the segment CF to the compression phase and the segment FA to the delivery phase. The position of the end compression point F is determined by the quantity of injected fluid and therefore both by the dimension of the entrance sections 25A and 25B and exit sections 27A and 27B and by the distribution pressure P2. Therefore, by acting on these parameters it is quite easy to vary the ratio between the flow rate sucked by the compressor at the pressure P1 and that of fluid injected at pressure P2 so as to optimize both the increase in the refrigerating capacity and the energy efficiency of the refrigerating machine on which the compressor with the distributor is installed.

FIG. 11 schematically shows a refrigerating circuit using a compressor of the type described above. FIG. 12 shows the respective transformation cycle of the coolant on a pressure-enthalpy diagram. The illustrated example refers to a cycle with CO₂ as coolant, but it should be understood that other adequate refrigerating fluids can be used. It is clearly apparent that the numerical values (of temperature and pressure) indicated below are indicated just by way of example, as they can vary according to the type of coolant R used and to the desired use conditions.

The refrigerating circuit comprises the compressor 1 supplying the coolant R at a pressure P3 of about 90 bar and temperature of about 85° C. towards a main heat exchanger X1 (through the point I), which is in turn connected (point II) to an exchanger-economizer X2 for cooling the fluid R. The fluid R passes (point III) from the exchanger X2 to the expansion valve X3, where the pressure is reduced to the pressure PI at about 25 bar, and it is subsequently supplied (point IV) to an evaporator X4 for evaporating the remaining part of the fluid still in the liquid state. The fluid in the form of steam at pressure P1 of about 25 bar is supplied (point V) from the evaporator X4 to the entrance 1L of the compressor 1.

According to a particularly advantageous embodiment of the present invention, an economizer circuit or auxiliary circuit is connected downstream of the condenser X1 (in the point II) to deviate part of the fluid R towards a secondary expansion valve X5 designed for decreasing the pressure P2 to about 50 bar. From here, the coolant R is supplied (point VI) to the exchanger-economizer X2, where it is heated and then completely evaporated. The main flow of the coolant flowing in the other branch of the exchanger-economizer X2 is cooled between the point II and the point III. The fluid R in the auxiliary circuit is subsequently supplied (point VII) to the entrance 14C of the distribution chamber 14A of the compressor 1 at a pressure P2 at about 50 bar. The fluid R coming from the suction 1L is compressed in the compression chamber and at the same time mixed with the fluid R coming from the connection 14C on the auxiliary circuit (point VIII). The fluid R is then further compressed, until it achieves the pressure P3 (point I).

The distributor 19 provided inside the compressor 1 supplies the fluid R at 50 bar in the compression chamber or chambers 5, 7 according to a preset phase, as described above.

FIG. 12 shows an usual semi-logarithmic diagram enthalpy-pressure, on which is represented the transformation cycle to which the fluid R is subject in the plant of FIG. 11. In the diagram of FIG. 12 on the axis of abscissas is shown the heat content of the fluid (the enthalpy) in kJ/kg and on the axis of ordinates the values of absolute pressure in bar. Briefly, this diagram is conventionally divided into three areas by a saturation bell: an area of undercooled liquid L (on the left of the bell), an area of overheated steam G (on the right of the bell) and an area L+G, where steam and liquid coexist in different percentages (within the bell).

On the curve representing the cycle the points I to VIII of the circuit of FIG. 11 are indicated. From this diagram it should be noted in particular that during the compression phase from the point V to the point I, passing through the point VIII, the injection of further fluid R at a pressure P2 entails a decrease in the temperature at the end of the compression (point I) relative to a traditional compressor: the temperature decrease will be the greater the greater the ratio between the rate of injected fluid from the connection 14C and that sucked from the connection 1L. Furthermore, this diagram shows how the heating of the fluid R in the main circuit from the point II to the point III involves an increase in the refrigerating effect.

FIG. 13 schematically shows another type of refrigerating plant using advantageously a compressor of the type described above. FIG. 14 shows the respective transformation cycle of the coolant on a pressure-enthalpy diagram.

Unlike that of FIG. 11, in the plant of FIG. 13 the pick-up point of the coolant R at the secondary expansion valve X5 is performed downstream of the exchanger-economizer X2 (point III) and therefore with a lower enthalpy content, as the coolant R has been already undercooled in the exchanger-economizer X2. The coolant passes from the point III in the secondary expansion valve X5, where the pressure is reduced to the pressure P2 (point IV). From here, the fluid R is supplied to the exchanger-economizer X2, where it is heated and completely evaporated (point VII), and conveyed to the entrance 14 C of the distribution chamber 14A of the compressor 1. The description of the diagram in FIG. 14 is completely similar to that of FIG. 12, except for the pick-up point of the coolant R at the secondary circuit, and it is therefore omitted for the sake of conciseness.

While specific embodiments of the invention have been shown and described in detail to illustrate the application of the principles of the invention, it will be understood that the invention may be embodied otherwise without departing from such principles.

Claims

1. A reciprocating compressor for a coolant, comprising:

at least one piston slidable in a compression chamber and actuated by a drive shaft, wherein at least one supply port exits in said at least one compression chamber a supply port for supplying a coolant at a distribution pressure intermediate between a delivery pressure and a suction pressure of the compressor, and to said at least one supply port a distribution system is associated, said distribution system comprising a distributor slidable in a distribution chamber in synchronous manner with said at least one piston, said distributor controlling opening and closing of said supply port in said compression chamber, said distributor being actuated by said at least one drive shaft through a connecting rod-crank mechanism.

2. A compressor as claimed in claim 1, wherein said distributor presents at least one passage duct to put in fluid connection said distribution chamber with said at compression chamber through said supply port.

3)-4). (canceled)

5. A reciprocating compressor as claimed in claim 1, further comprising an inlet system to supply the coolant in said distribution chamber at the distribution pressure intermediate between the suction pressure and the delivery pressure of the compressor.

6. A reciprocating compressor as claimed claim 1, wherein said distributor is designed to open and to close gradually an entrance end of said at least one supply port, an exit end of said supply port being opened and closed by said at least one piston.

7. A reciprocating compressor as claimed in claim 1, wherein said distributor performs a shorter stroke relative to that of said at least one piston.

8. A reciprocating compressor as claimed in claim 2, wherein said at least one passage duct presents at least one entrance aperture on a head of the distributor and at least one exit aperture arranged in an intermediate position on a lateral shell of said distributor.

9. A reciprocating compressor as claimed in claim 8, wherein said at least one exit aperture exits in a compartment obtained on said lateral shell of the distributor, in said compartment being inserted and slidable a shaped seal sliding block, which presents a through channel close said at least one exit aperture, said seal sliding block being designed to be pushed by the pressure of the coolant against a wall of said distribution chamber.

10. A reciprocating compressor as claimed in claim 9, wherein said distributor comprises at least a first seal between the distribution chamber and the seal line block and at least one second seal between said seal sliding block and one compartment of a casing of the compressor.

11. A reciprocating compressor as claimed in claim 1, wherein said at distributor and said at least one piston are mutually phased so as to open said at least one supply port when said at least one piston is about at a bottom dead center and/or said at least one piston is in compression phase.

12. A reciprocating compressor as claimed in claim 1, further comprising another piston to provide at least two compression pistons slidable in respective compression chambers, each of said compression chambers presenting at least one supply port for a fluid connection with said distribution system.

13. A reciprocating compressor as claimed in claim 2, wherein said at least one passage duct of the distributor is shaped so as to supply fluid in said at least two compression chambers in different positions of the drive shaft.

14. A reciprocating compressor as claimed in claim 1, further comprising another piston to provide at least one pair of compression pistons displaced from each other by a phase angle and said distributor is displaced relative to said at least one pair of said compression pistons by respective intermediate angles equal to about half of said phase angle.

15. A reciprocating compressor as claimed in claim 8, wherein said at least one piston presents at least one first seal close to the head, and at least one seal close to an end opposite to the head.

16. A refrigerating plant comprising:

a main circuit for a coolant, with a condenser, an expansion member, an evaporator, and a compressor comprising at least one piston slidable in a compression chamber and actuated by a drive shaft, wherein in said at least one compression chamber a supply port exits for supplying a coolant at a distribution pressure intermediate between a delivery pressure and a suction pressure of the compressor, and to said at least one supply port a distribution system is associated, said distribution system comprising a distributor slidable in a distribution chamber in synchronous manner with said at least one piston, said distributor controlling opening and closing of said supply port in said compression chamber, said at least one distributor being actuated by said at least one drive shaft through a connecting rod-crank mechanism;

an auxiliary circuit with an entrance of coolant cooled and/or condensed at high pressure, and an exit connected to said compressor, between said entrance and said exit being arranged an expansion member and an exchanger economizer, through which the coolant circulating in the auxiliary circuit cools a flow of the coolant circulating in the main circuit, the fluid at the exit of the auxiliary circuit presenting a pressure intermediate between the suction pressure and the delivery pressure of the compressor and being inserted, through said distribution system in said at least one compression chamber of the compressor.

17. A refrigerating plant comprising:

a main circuit for a coolant, with a condenser, an expansion member, an evaporator, and a compressor comprising at least one piston slidable in a compression chamber and actuated by a drive shaft, wherein in said at least one compression chamber a supply port exits for supplying a coolant at a distribution pressure intermediate between a delivery pressure and a suction pressure of the compressor, and to said at least one supply port a distribution system is associated, said distribution system comprising a distributor slidable in a distribution chamber in synchronous manner with said at least one piston, said distributor controlling opening and closing of said supply port in said compression chamber, said at least one distributor being actuated by said at least one drive shaft through a connecting rod-crank mechanism;

an auxiliary circuit with an entrance of undercooled coolant, and an exit connected to said compressor, between said entrance and said exit being arranged an expansion member and an exchanger economizer, through which the coolant circulating in the auxiliary circuit cools a flow of the coolant circulating in the main circuit; the fluid at the exit of the auxiliary circuit presenting the distribution pressure intermediate between the suction pressure and the delivery pressure of the compressor and being inserted, through said distribution system in said at least one compression chamber of the compressor.