EP1886075A1

EP1886075A1 - Refrigeration plant

Info

Publication number: EP1886075A1
Application number: EP06701814A
Authority: EP
Inventors: Hermann Renz; Günter DITTRICH
Original assignee: Bitzer Kuehlmaschinenbau GmbH and Co KG
Current assignee: Bitzer Kuehlmaschinenbau GmbH and Co KG
Priority date: 2005-02-17
Filing date: 2006-01-24
Publication date: 2008-02-13
Anticipated expiration: 2026-01-24
Also published as: DE102005009173A1; US20080011014A1; US7451617B2; EP1886075B1; CN100538206C; CN101120213A; WO2006087075A1

Abstract

According to the invention, a refrigeration plant comprising a refrigerant circuit (10), in which a main mass flow (56) of a refrigerant flows which may be optimally matched to differing operating conditions may be achieved, whereby at least two refrigerant compressors (12a, 12b, 12c) are arranged in the refrigerant circuit which may be individually switched on for compression of the main mass flow and at least two of the refrigerant compressors have an additional compressor stage (70). Each of the compressor stages may be optionally applied to compression of refrigerant from the main mass flow or for compression of refrigerant from the additional mass flow (86) and a controller (90) is provided which, in a first operating mode depending on operating conditions, such a number of additional compressor stages may be switched on for compression of refrigerant from the additional mass flow that the expansion cooling device (24, 28) condenses the main mass flow and reduces the enthalpy thereof.

Description

refrigeration plant

The invention relates to a refrigeration system comprising a refrigerant circuit in which a main mass flow of a refrigerant - preferably carbon dioxide - is guided, arranged in the refrigerant circuit high-pressure side heat exchanger, a refrigerant circuit arranged in the expansion kühieinrichtung that cools the main mass flow of the refrigerant in the active state and thereby generating an additional mass flow of gaseous refrigerant, a reservoir arranged in the refrigerant circuit for liquefied refrigerant, at least one expansion unit for liquefied refrigerant of the main mass flow arranged in the refrigerant circuit, which provides an expansion element and a downstream low-pressure side cooling capacity, ie the enthalpy of refrigeration. by means of elevating, having heat exchanger, and at least one arranged in the refrigerant circuit refrigerant compressor, which is a main compressor stage and at least one additional compressor stage driven jointly with the main compressor stage, which compresses both refrigerants to high pressure, wherein the main compressor stage and the at least one additional compressor stage are usable so that either the main compressor stage refrigerant from the main mass flow and the additional compressor stage refrigerant from the additional mass flow or the main compressor stage and the Additional compressor stage Compacting refrigerant from the main mass flow.

Such refrigeration systems are known from the prior art, which are designed for conventional refrigerants.

For example, such refrigeration systems are described in EP 0 180 904 A2. Based on this known prior art, the invention has the object to provide a refrigeration system that can be optimally adapted to different operating conditions.

This object is achieved in a refrigeration system of the type described above according to the invention in that in the refrigerant circuit at least two refrigerant compressors are arranged, which are individually switchable for compressing the main mass flow, that at least two of the refrigerant compressor each have at least one additional compressor stage that each of the additional compressor stages is selectively usable for compressing refrigerant from the main mass flow or for compressing refrigerant from the additional mass flow and that a control is provided with which in a first operating mode depending on operating conditions such a number of additional compressor stages for compressing refrigerant from the additional mass flow can be switched that the expansion cooling device liquefies the main mass flow and reduces its enthalpy.

The advantage of the solution according to the invention lies in the fact that, due to the variable connectability of the additional compressor stages, this makes it possible to adapt the liquefaction of the main mass flow and the enthalpy reduction to the various operating conditions and thus to always keep it in an optimum range.

It is particularly favorable if the expansion cooling device reduces the enthalpy of the main mass flow by at least 10%. It is even more advantageous if the expansion cooling device reduces the enthalpy of the main mass flow by at least 20%. The refrigeration plant can be used particularly advantageously when the first operating mode corresponds to a critical operation, for example with carbon dioxide as the refrigerant

Under a supercritical operation is to be understood that the high-pressure refrigerant compressed in the high-pressure side heat exchanger can not be cooled to a temperature corresponding to a boiling line and saturation curve of the refrigerant isotherms, but can only be cooled to a temperature that a outside the boiling line and saturation curve extending isotherms, so that it does not come to a liquefaction of the refrigerant.

Furthermore, a particularly favorable embodiment provides that the expansion cooling device converts the main mass flow into a thermodynamic state whose pressure and enthalpy are lower than pressure and enthalpy of a maximum of the saturation curve or boiling line in an enthalpy / pressure diagram.

Preferably, the thermodynamic state of the main mass flow caused by the expansion cooling device is close to the boiling point of the enthalpy / pressure diagram, in particular substantially at the boiling point or at an enthalpy which is lower than the enthalpy corresponding to the boiling curve at the respective pressure.

The expansion cooling device can basically be designed in any desired manner. However, a particularly favorable solution provides that the expansion cooling device has an expansion valve for expanding refrigerant to an intermediate pressure and that the intermediate pressure of Exapansionskühlein- direction is adjustable by adding the appropriate number of additional compressor stages.

Furthermore, for example, the expansion cooling device could work so that only an expansion of the additional mass flow forming refrigerant takes place.

However, it is even more advantageous if the expansion valve of the expansion cooling device expands the refrigerant of the main mass flow and of the additional mass flow to the intermediate pressure.

With regard to the arrangement of the reservoir for the liquid refrigerant, no details have been given so far. Thus, a particularly favorable solution provides that the expansion cooling device includes the reservoir for the liquid refrigerant of the main mass flow and thus simplifies the construction of the refrigeration system according to the invention.

A structurally particularly preferred solution provides that the expansion valve converts the expanded refrigerant from the main mass flow and the additional mass flow into a container in which the reservoir for the liquid refrigerant of the main mass flow forms over which there is a vapor space, from which then the additional mass flow forming Refrigerant is removed, so that a portion of the refrigerant evaporates and thereby cools the main mass flow or even subcooled. A further advantageous embodiment of the refrigeration system according to the invention provides that in a second operating mode, the expansion cooling device is in the inactive state and does not cause cooling of the main mass flow.

That is, in this case, no additional mass flow of refrigerant is produced and thus the refrigeration system according to the invention can be operated in the conventional manner known manner by a circuit of the entire refrigerant in the form of the main mass flow.

It is expediently provided in such a second operating mode of the refrigeration plant that all additional compressor stages compress refrigerant of the main mass flow.

Furthermore, it is provided in a second operating mode of the refrigeration system according to the invention that the reservoir for liquid refrigerant of the main mass flow is under high pressure.

In particular, it is provided in one embodiment that the second operating mode corresponds to a subcritical operation of the refrigeration system.

Under a subcritical operation of the refrigeration system in the sense of the solution according to the invention is to be understood that in the high-pressure side heat exchanger such a strong cooling of the high-pressure compressed refrigerant is possible that this passes into a thermodynamic state, which is below the saturation curve or boiling line, that is in the Area of coexistence of liquid and vapor, is, and is thus cooled so that a liquefaction of the refrigerant takes place through the high-pressure side heat exchanger.

In order to operate the refrigeration system according to the invention always under optimal conditions, in particular in adaptation to the required cooling capacity, it is provided that the controller controls the refrigerant compressor according to the required cooling capacity, that is, the refrigerant compressor can be operated either with variable speed and / or can be switched on or off.

It is particularly useful in this case if the controller is able to individually connect or disconnect the refrigerant compressor according to the required cooling capacity, that is, by Einzelzu- or shutdown of the at least two refrigerant compressor in the refrigerant circuit is possible, the compressor power to the to adjust the required cooling capacity and thus always operate the refrigeration system according to the invention optimally.

With regard to the design of the additional compressor stages in relation to the respective main compressor stage so far no further details have been made. So it is particularly advantageous if each refrigerant compressor is dimensioned with additional compressor stage so that the mass flow of refrigerant of the additional mass flow compressed by the additional compressor stage corresponds maximally to the compressed mass flow of refrigerant of the main mass flow in this refrigerant compressor from the main compressor stage. Furthermore, the possibilities given by the control of adjusting the additional mass flow and the intermediate pressure can advantageously be exploited in that the refrigerant compressors with additional compressor stage are dimensioned such that the additional compressor stages of different refrigerant compressors compress different mass flows of refrigerant of the additional mass flow.

Thus, a suitable variation of the additional mass flow to be compressed can be achieved by suitably selecting the additional compressor stages provided for compressing refrigerant of the additional mass flow, in particular by suitable combination for different mass flows of additional compressor stages, without the power of the main compressor stages having to be changed for this purpose.

Since in refrigeration systems which are to be operated in the supercritical range, a very high pressure difference must be generated during the compression of the refrigerant, it is preferably provided that the refrigerant compressors with additional compressor stage are reciprocating compressors.

In such reciprocating compressors, each of the refrigerant compressor with additional compressor stage is expediently designed so that it has at least one cylinder for the additional compressor stage and at least one cylinder for the main compressor stage.

Such a refrigeration system can be realized in a particularly favorable manner if, for each refrigerant compressor with additional compressor stage, the number of cylinders for the main compressor stage is greater than the number of cylinders for the additional compressor stage. Furthermore, a solution of the refrigeration system according to the invention that is particularly favorable with regard to the variable adjustability of the additional mass flow provides that the additional compressor stages of different refrigerant compressors have a different stroke volume from the refrigerant compressors with additional compressor stage, thereby resulting in a particularly wide range of stroke volumes even in a different combination of additional compressor stages the additional mass flow is available for selection.

Furthermore, another solution which is suitable with regard to its variability provides that the ratio of the stroke volume of the additional compressor stage to the displacement volume of the main compressor stage is different with respect to at least one of the other refrigerant compressors with additional compressor stage for each refrigerant compressor with additional compressor stage, so that not only the stroke volumes of the additional compressor stages can be summarized by a suitable selection and combination with each other to the largest possible variation bandwidth, but also the stroke volumes of the main compressor stages.

A further advantageous embodiment of the refrigeration system according to the invention provides that in the first operating mode, the reservoir for liquefied refrigerant operates at an intermediate pressure and between the high-pressure side, the refrigerant cooling heat exchanger and the reservoir for liquefied refrigerant an additional expansion unit with an expansion element and a downstream cooling capacity Providing heat exchanger is provided. With this additional expansion unit, the thermodynamic efficiency of the refrigeration system according to the invention can be further improved, since the evaporation temperature in this Additional expansion unit is higher, which requires that the provided cooling capacity at a higher temperature level, for example, for room cooling or room air conditioning, can be used.

In particular, in all the above exemplary embodiments, a significantly improved thermodynamic efficiency can be achieved under supercritical operating conditions. In addition, the cooling capacity is higher at a defined compressor stroke volume and the power characteristic in relation to the ambient temperature flatter, which has a positive effect on the control characteristics of the refrigeration system.

The higher efficiency in supercritical operation is due, in particular, to the fact that the evaporation of the additional mass flow takes place at a higher pressure level than the evaporation in the suction-side heat exchangers of the expansion units. This leads to an improvement of the thermodynamic efficiency with the result of a reduced energy requirement for a defined cooling capacity.

In particular, by the expansion of the main mass flow and the additional mass flow in conjunction with the suction of the additional mass flow cooling of the refrigerant of the main mass flow at saturation pressure to the boiling or saturation curve. This increases the enthalpy difference for evaporation and overheating. The increase in the enthalpy difference is higher in percentage than the proportion of the compressor power that must be expended for compressing the additional mass flow. In addition to the aforementioned improvement in efficiency, this also leads to a higher cooling capacity - based on an identical total displacement volume of the refrigeration system. Furthermore, it is advantageous in the refrigeration system according to the invention, in particular when carbon dioxide is used as the refrigerant, if the refrigerant compressors have cylinder heads in which inlet chambers and outlet chambers are substantially thermally decoupled, so that the heating of the refrigerant during compression to high pressure and the associated heating of the outlet chambers is substantially no heating of the inlet chambers with the refrigerant to be sucked and thus no negative effect on the compressor performance.

To be able to use the additional compressor stages either for compressing the main mass flow or for compressing the additional mass flow, it is basically conceivable to provide controlled valves which supply the additional compressor stages either with refrigerant of the main mass flow or of the additional mass flow.

However, a structurally particularly simple solution provides that a check valve is provided for connecting an inlet chamber of the additional compressor stage to the low-pressure connection of the main compressor stage, so that inevitably, when the additional mass flow is interrupted, the additional compressor stage compresses refrigerant of the main mass flow.

A particularly simple solution provides that the check valve connects the inlet chamber of the additional compressor stage with the inlet chamber of the main compressor stage. Another advantageous solution provides that the check valve is provided in a valve plate of the respective refrigerant compressor. This solution has the advantage that the already equipped with valves valve plate only needs to be provided with an additional check valve and thus the check valve can be very easily mounted.

It is especially expedient if a connecting channel between the low-pressure connection and the check valve runs in a cylinder housing and can be molded into it in the same way as the inlet channel for supplying the main compressor stage with refrigerant supplied via the low-pressure connection.

Further features and advantages of the invention are the subject of the following description and the drawings of some embodiments.

In the drawing show:

1 shows a schematic representation of a first embodiment of a refrigeration system according to the invention;

FIG. 2 shows a schematic representation of one of the refrigerant compressor with main compressor stage and additional compressor stage used in the refrigeration system according to the invention in the first exemplary embodiment; FIG. 3 shows a representation of the pressure [P] versus the enthalpy [h] in the case of a subcritical cycle process that can be implemented in the first exemplary embodiment and a possible supercritical cycle process that does not correspond to the invention;

4 shows an illustration of the pressure [P] versus the enthalpy [h] in a supercritical cycle according to the invention with expansion of the high-pressure compressed refrigerant to an intermediate pressure and simultaneous reduction of the enthalpy by suction an additional mass flow;

Fig. 5 is a schematic representation of a refrigerant compressor in a second embodiment of the invention

Cooling system;

Fig. 6 is a schematic representation of a third embodiment of a refrigeration system according to the invention;

7 is a perspective view of a cylinder head of a first preferred embodiment of a refrigerant compressor for a refrigeration system according to the invention;

8 shows a perspective view of the cylinder head according to FIG. 7 with its underside pointing upwards; 9 shows a partial section through a second preferred embodiment of a refrigerant compressor for the refrigeration system according to the invention and

10 is a perspective view of a valve plate of the second preferred embodiment of the refrigerant compressor of FIG. 9.

An exemplary embodiment of a refrigeration system illustrated in FIG. 1 comprises a refrigerant circuit designated as a whole by 10, in which several, for example three, refrigerant compressors 12a to 12c are arranged whose high-pressure connections 14a to c are connected to a high-pressure line 16 of the refrigerant circuit 10.

The high-pressure line 16 leads to a high-pressure side heat exchanger 18 which cools the refrigerant compressed to high pressure PH, for example with a flow 20 of a cooling medium, the cooling medium preferably being ambient air flowing through the heat exchanger 18.

But it is also conceivable, instead of ambient air, another cooling medium, for example, water or the like to provide.

From the heat exchanger 18, a further high-pressure line 22 leads to an expansion valve 24 and to a bypass valve 26 connected in parallel to the expansion valve 24, both of which open into a container 28 which is designed to include a reservoir 30 for liquid refrigerant which is always a volume 32 of liquid refrigerant is present, which - as described in detail below - represents a buffer volume for liquid refrigerant in the refrigerant circuit 10.

From the reservoir 30, a conduit 34 leads to expansion units 40, for example four expansion units 40a to 40d connected in parallel.

The line 34 is connected to the reservoir 30 in such a way that it essentially only leads liquid refrigerant to the expansion units 40 and thus can be operated and in particular regulated in a known manner, since there is always an expansion of liquid refrigerant, essentially without Gas content, takes place.

The control of expansion units 40, which are supplied with liquid refrigerant, corresponds to the type of control in known refrigeration systems.

Each of the expansion units 40 includes a shut-off valve 42, an expansion valve 44 that expands the liquid refrigerant, and a low-pressure side heat exchanger 46 that is capable of discharging cooling power due to the expanded refrigerant as indicated by the arrow 48.

The heat exchangers 46 of the parallel-connected expansion units 40 are connected to a common low-pressure line 50, which leads to low-pressure connections 52a to 52c of the refrigerant compressor 12a to 12c.

The sum of all the partial mass flows 54a, 54b, 54c and 54d of the refrigerant that are passing through the expansion units 40 and that are collected by the low-pressure line 50 form a main mass flow 56 of the Refrigerant circuit 10, which in turn is divided again into partial mass flows 58 a, 58 b and 58 c, which are sucked by the refrigerant compressors 12 a to 12 c via the low-pressure ports 52 a to 52 c and compressed to high pressure PH, through the high-pressure ports 14 a to 14 c of the refrigerant compressor 12 again withdraw.

Since the partial mass flows 54a to 54d have been taken out of the line 34, the line 34 is also flowed through by the main mass flow 56 following the reservoir 30, which then divides again onto the partial mass flows 54a to 54d.

As shown in FIG. 2, for example, each of the refrigerant compressors 12 is formed as a reciprocating compressor and includes a cylinder housing 60 in which, for example, four cylinders 62a to 62d are provided in which refrigerant can be compressed by oscillatingly moving pistons 64a to d.

In a refrigerant compressor 12 configured in accordance with the invention, not all the cylinders 62a to 62d now operate as a single compressor stage, but for example the cylinders 62a to 62c are combined to form a main compressor stage 66 in which these three cylinders 62a to 62c operate in parallel, ie all three cylinders 62a to 62c suck in refrigerant via the respective low-pressure port 52 and deliver refrigerant compressed to high-pressure PH to the respective high-pressure port 14. Furthermore, the cylinder 62d, which is driven together with the other cylinders of the main compressor stage 66 and in the same way as these by a drive motor 68, is operated as a separate auxiliary compressor stage 70, which is also connected to the high-pressure connection 14 on the output side, but is able to either to suck in refrigerant via an additional connection 72 or via the low-pressure connection 52.

In the simplest case, a check valve 76 is provided in the connecting channel 74 running between the additional connection 72 and the low-pressure connection 52, which blocks the connection channel 74 when the pressure at the additional connection 72 is higher than that at the low-pressure connection 52, so that always, when refrigerant is present at the additional port 72 at a higher pressure than at the low pressure port 52, the connecting channel 74 is blocked and thus the additional compressor stage 70 sucks refrigerant via the auxiliary port 72. But it can also be provided a controlled valve.

If, however, the additional connection 72 is shut off or blocked so that no refrigerant can be sucked in through it, then the check valve 76 opens and the additional compressor stage 70 sucks in refrigerant via the low-pressure connection 52 and compresses it to high pressure PH, in the same way as the main compressor stage 66.

As shown in FIG. 1, the auxiliary ports 72a to 72c of the refrigerant compressors 12a to 12c are respectively connected via shutoff valves 80a to 80c to a distribution pipe 82 which open into the tank 28 so as to be capable of remove refrigerant vaporized from a vapor space 84 of the container 28. The vaporized refrigerant discharged from the container 28 from the distribution line 82 forms an additional mass flow 86 which can be distributed from the distribution line 82 to the additional compressor stages 70 in order to be compressed by the same to the high pressure PH.

The additional mass flow 86 can thus be controlled by opening or closing individual ones of the shut-off valves 80a to 80c.

To control the refrigeration system is a total of 90 designated

Control provided, which is able to individually control the individual shut-off valves 80a to 80c.

If all the shut-off valves 80a to 80c are closed, no additional mass flow 86 flows through the distribution line 82 and no additional mass flow 86 is compressed in the additional compressor stages 70, so that only the main mass flow 56 is compressed in the refrigerant circuit 10 with all the cylinders 62 and is expanded.

However, if one of the shut-off valves 80a to 80c or all shut-off valves 80a to 80c are opened, the additional mass flow 86 flows through the distribution line 82 is supplied to the additional compressor stages 70, which are connected via the open shut-off valves 80a to 80c with the distribution line 82, and thus is compressed by the corresponding additional compressor stages 70 of the respective refrigerant compressor 12, so that in addition to the main mass flow 56 of the additional mass flow 86 flows through both the high pressure line 16 and through the high-pressure side heat exchanger 18 and the other High pressure line 22 is supplied to the container 28, wherein in the container 28 is a separation between the main mass flow 56 and the additional mass flow 86 to the effect that the main mass flow 56 is supplied via the line 34 to the expansion units 40, while the additional mass flow 86 via the distribution line 82 the corresponding Additional compressor stages 70 is supplied and thus does not flow through the expansion units 40.

A refrigeration system formed in this way can be operated as a refrigerant in particular with carbon dioxide (CO ₂ ) as follows:

If a sufficiently strong cooling of the refrigerant compressed to high pressure PH in the high-pressure-side heat exchanger 18 is possible, then the refrigeration system can be operated in the so-called subcritical cycle. With carbon dioxide as the refrigerant, this presupposes that the temperature of the cooling medium 20 supplied to the high-pressure side heat exchanger 18 is on the order of about 23 ° C. or below. In this case, cooling of the refrigerant compressed to high pressure PH results in liquefaction thereof, so that the bypass valve 26 is opened by the controller 90 and the liquid refrigerant from the further high-pressure line 22 is directly supplied to the liquid refrigerant reservoir 30.

This liquid refrigerant then forms the main mass flow 56, which is distributed via the line 34 to the individual expansion units 40, provided that these are switched on by the controller 90, that is, the shut-off valves 42a to d are open. The activation of the individual expansion units 40a to d takes place depending on whether refrigeration capacity 48 is to be made available in the region of the respective low-pressure-side heat exchanger 46 or not.

The refrigerant expanded in the individual expansion units 40a to 40d is then supplied via the low-pressure line 50 to the individual low-pressure connections 52a to 52c of the individual refrigerant compressors 12a to c.

The controller 90 does not necessarily operate all the refrigerant compressors 12a to 12c in the full load range, but can operate either individual of the refrigerant compressors 12a to 12c in the full load range or individual or all refrigerant compressors 12a to 12c in the partial load range, ie with reduced rotational speed of the respective drive motor 68. However, it is also possible, on the part of the controller 90, to completely shut off the individual refrigerant compressors 12a to 12c, for example when only a part of the expansion units 40a to 40d is to be provided with refrigerating capacity at their respective heat exchanger 46.

In addition, the controller 90 closes the shut-off valves 80a to 80c in subcritical operation, so that in all refrigerant compressors 12a to 12c the additional compressor stages 70 suck refrigerant from the main mass flow 56 via the respective check valve 76 and compress it to high pressure PH.

Such a cycle for the subcritical operation is shown in Fig. 3 by the broken lines, wherein the state in point A represents the incipient compression of refrigerant from the main mass flow 56 through the respective refrigerant compressor 12, which is completed in the state in point B. Starting from the state in point B, the refrigerant compressed under high pressure PH is cooled down to a state at point C which is approximately at the saturation curve or boiling line 96 for carbon dioxide as the refrigerant.

This now cooled, but liquefied in the heat exchanger 18 refrigerant can now be supplied to the individual expansion units 40 in this state, with the expansion valve 44 each of the expansion units 40 isenthalpe relaxation of the refrigerant takes place, resulting in a reduction of pressure associated with a reduction in temperature leads, so that the state is reached in point D in Fig. 3.

Starting from the state in point D, the cooling capacity 48 can now be made available in the respective low-pressure side heat exchanger 46 by enthalpy increase, until the state in point A is reached, which represents the refrigerant in terms of enthalpy and pressure, which via the low-pressure line 50th the low pressure ports 52 of the refrigerant compressor 12 is supplied.

However, if no cooling medium is available which is able to cool the refrigerant to a temperature of the order of magnitude of 23 ° C., only one cooling medium is available, which merely cools to relatively high temperatures of the refrigerant, for example above 31 ° C. allowed, so with open bypass valve 26 and ineffective expansion valve 24 with the refrigeration system shown in FIG. 1, only a so-called supercritical cycle (drawn in Fig. 3 drawn) possible, in which one the refrigerant would have had to compress to a higher pressure in accordance with the condition at point B ¹ in FIG. 3, wherein a subsequent cooling in the high-pressure side heat exchanger 18 leads to a state in point C in FIG. 3, which lies outside the saturation curve 96.

The refrigerant in the state in point C in Fig. 3 is still gaseous. A subsequent isenthalp relaxation of the refrigerant in the individual expansion units 40 would then lead to the state in point D 'in Fig. 3, the consequence would be that the expansion valves 44 of the expansion units 40 gaseous refrigerant would be supplied and gaseous refrigerant would have to be expanded. Such expansion of a gaseous refrigerant is subject to other control characteristics, so that thus the previously known control characteristics for the expansion valves 44 are not suitable.

For this reason, in supercritical operation of the refrigeration system, no supercritical cycle shown in FIG. 3 occurs from point A to B ¹ to C and D 'and then back to A, but the bypass valve 26 is closed by the controller 90 and the expansion valve 24 becomes activated, so that the refrigerant entering from the further high-pressure line 22 into the container 28 can be expanded by the expansion valve 24 to an intermediate pressure PZ corresponding to the state Z in Fig. 4. In addition, at the intermediate pressure PZ by evaporating the refrigerant, the temperature can be lowered so far, and thus also the enthalpy, that there is liquid refrigerant of the main flow 56 in the container 28, whose state corresponds to the state at point C on the boiling line 96 in FIG. 4 corresponds. In order to be able to move from the state in point C to the state in point C of FIG. 4, it is necessary to prescribe the intermediate pressure PZ in the container 28 and to stabilize this intermediate pressure PZ by sucking off the vaporized refrigerant, the evaporated refrigerant containing the additional mass flow 86 results, which must be removed from the vapor space 84 of the container in order to maintain the intermediate pressure PZ at the desired level can.

The state in point C is at a compared to a maximum 98 of the boiling line 96 by more than 20% lower value of the enthalpy [h] which is achieved by the evaporation of the additional mass flow forming refrigerant, the state in point C in Fig. 4 either essentially on the boiling line 96 or optionally with additional cooling, eg is located at a slightly lower enthalpy than the enthalpy of the state in point C via a heat exchanger penetrated by the expanded main mass flow.

For this purpose, the controller 90 must open at least part of the shut-off valves 80a to 80c or all shut-off valves 80a to 80c, thereby causing refrigerant from the additional stream 86 to be sucked in by the additional compressor stages 70 to maintain the intermediate pressure PZ in the container 28 and compressed to high pressure PH becomes.

Thus, the refrigerant of the main mass flow 56 can be supplied via the line 34 to the individual expansion units 42a to 42c and transferred by isenthalp relaxation in the expansion units 40 by means of the expansion valves 44 in the designated in Fig. 4 with point D state, in which enthalpy increase up to State at point A in the 3, it can be seen from the comparison with FIG. 3 that the cooling capacity provided is greater than in the case of a supercritical cyclic process corresponding to the states in the points A, B ¹ , C, D ¹ in Rg. 3.

The advantage of the inventive concept can be seen in the fact that opens the possibility optimally to choose the high pressure PH according to the course of the isotherms of the refrigerant used, without looking back on the downstream expansions must be taken.

Furthermore, the intermediate pressure PZ can also be optimized by suitable variation of the additional mass flow, in such a way that the decrease of the enthalpy of the main mass flow is higher than the percentage of the delivery volume of the total delivery volume of the compressor required for the additional mass flow, so that the by compressing the additional mass flow conditional loss of delivery volume is overcompensated by the decrease in the enthalpy of the main mass flow.

The cyclic process for maintaining the intermediate pressure PZ by compressing the additional mass flow 86 is shown in dotted lines in FIG. 4 and proceeds from the state at point Z by enthalpy increase of the vaporized refrigerant to the state at point A "and from state at point A" to state at Point B ", which in turn is at the high pressure PH, and from the state in point B" to the state in point C and from the state in point C to the state in point Z. In such a supercritical, shown in Fig. 4 cycle between the states in the points A, B ', C, C and D, if the intermediate pressure PZ is to be set optimally, the resulting additional mass flow 86 in relation to the main mass flow 56 is not constant but varies depending on how many expansion units 40 are activated in the refrigerant circuit 10 and how high the temperature of the cooling medium 20 supplied to the high-pressure side heat exchanger 18 is.

In order to be able to compress an additional intermediate pressure PZ, which is maintained according to optimized operating conditions, over the additional compressor stages 70, the additional compressor stages 70 of the refrigerant compressors 12 are designed in such a way that at maximum cooling capacity to be delivered by all expansion units 40 and maximum temperature of the cooling medium 20 is still an optimized supercritical operation is possible and the resulting additional mass flow 86 can be compressed to maintain a suitable level of the intermediate pressure PZ of the totality of the active auxiliary compressor stages 70 to high pressure PH.

If, on the basis of this operating state, more favorable operating conditions prevail, the controller 90 can either reduce the speed of the drive motors 68 of one or more of the refrigerant compressors 12 or shut off one of the refrigerant compressors 12, thereby eliminating both the compressor capacity of the main compressor stage of this refrigerant compressor 12 and the compressor capacity the additional compressor stage 70. However, if the operating conditions change such that, for example, cooling medium 20 is available at a lower temperature, the additional mass flow 86 changes, since less refrigerant must be evaporated to liquid refrigerant in the state in point C of FIG. 4 at a suitable intermediate pressure PZ receive.

In this case, the controller 90 has the ability to adjust by closing one or two of the shut-off valves 80a to 80c, the compressor power of the additional compressor stages 70 to the lower required additional mass flow 86 and thus maintain an optimized intermediate pressure PZ in the container 28.

The additional compressor stages 70, in which the shut-off valves 80 have been closed, then suck in refrigerants of the corresponding low-pressure connection 52 and thus compress refrigerant from the respective main mass flow 56.

The inventive concept thus allows an optimal adaptation of the intermediate pressure PZ by adjusting the compressor power required for the compression of the additional mass flow 86 of the additional compressor stages 70a to 70c independently of the compressor capacity of the main compressor stages 66.

In principle, it would be possible to design the refrigerant compressors 12a to 12c identically so that each main compressor stage 66 and each additional compressor stage 70 can provide the same compressor performance. However, it is even more advantageous for adapting to different operating conditions when the refrigerant compressors 12a to 12c are formed such that a second one of the refrigerant compressors 12a to 12c has twice the compressor capacity of the first refrigerant compressor, and a third refrigerant compressor has twice the compressor capacity of the second refrigerant compressor the doubling of compressor performance relates to both the main compressor stages 66 and the auxiliary compressor stages 70.

This makes it possible to obtain different multiples of the compressor capacity of the first refrigerant compressor by different combinations of the first, second and third refrigerant compressor, for example, twice the compressor capacity of the first refrigerant compressor alone by operation of the second refrigerant compressor, the triple compressor capacity of the first refrigerant compressor by operation of the first and second refrigerant compressor, the four times the cooling capacity by operating the third refrigerant compressor and the fivefold cooling capacity by operating the third refrigerant compressor in combination with the first refrigerant compressor and the seven times the compressor capacity by combining the first, second and third refrigerant compressor.

With regard to the compressor power of the additional compressor stages 70, further variations are conceivable, namely in that when all three refrigerant compressors 12a to 12c are operated, the maximum power of the additional compressor stages 70 for the additional mass flow 86 is available, which is seven times the compressor capacity of the first refrigerant compressor equivalent. However, it is also conceivable to use any integer multiple of the compressor power of the additional compressor stage 70 of the first refrigerant compressor with analogous application of the procedure described above by opening shut-off valves 80 and connecting the individual additional compressor stages 70 of the individual refrigerant compressor 12 with the distribution line 82 for compressing the additional mass flow 86.

In a second exemplary embodiment of a refrigeration system according to the invention, illustrated in FIG. 5, the refrigerant compressors 12 'are designed, for example, such that they have two additional compressor stages 7O ₁ and 7O ₂ , each having its own additional connections 7O _x and 7O ₂ .

For example, the cylinder 62c, the additional compressor stage 7O ₂ and the cylinder 62d, the additional compressor stage 7O ₁ , while the cylinders 62a and 62b form the main compressor stage 66.

Such a construction of one of the refrigerant compressor 12 or all refrigerant compressor 12 provides even greater variability in terms of compressing the compressor power available for compressing the additional mass flow 86, since the individual additional compressor stages 7O ₁ and 7O ₂ individually or jointly either by opening the corresponding shut-off valves 80 with the Distribution line 82 can be connected or can be used to compress refrigerant of the main mass flow 56.

Incidentally, the second embodiment of the refrigeration system according to the invention corresponds to the first embodiment, so that the description of the first embodiment of the refrigeration system according to the invention can be fully incorporated by reference. Also, a third embodiment of the refrigeration system according to the invention, shown in Fig. 6, based on the first embodiment of the inventive refrigeration system, the same parts are provided with the same reference numerals, so that with respect to the description of the same fully incorporated by reference to the comments on the first embodiment ,

In the third embodiment, the bypass valve 26 and the expansion valve 24 are still a Zusatzzppansionseinheit 100 connected in parallel.

The additional expansion unit 100 in turn comprises a shut-off valve 102, an expansion valve 104 and a high-pressure side heat exchanger 106, from which cooling power characterized by an arrow 108 can be dissipated.

With this additional expansion unit is also possible to expand refrigerant from the high-pressure line 22 and thus obtain outward available cooling capacity 108, wherein the refrigerant is expanded only to the present in the container 28 intermediate pressure PZ.

This makes it possible to operate in supercritical operation additionally a working at a higher temperature level heat exchanger 106 and thereby increase the efficiency of the refrigeration system. However, the refrigerant which is expanded in the additional expansion unit 100 does not cause a cooling effect for the main mass flow 56 and has to be removed via the additional mass flow 86 and re-compressed by the additional compressor stages 70.

Incidentally, the third embodiment of the refrigeration system according to the invention works similar to the first embodiment, so that also with respect to the function is fully incorporated by reference to the first embodiment.

With regard to the structure of the refrigerant compressor so far no details have been made. For example, conventional reciprocating compressors can be used as refrigerant compressors.

It is particularly advantageous if, in a first preferred embodiment of such a refrigerant compressor, a cylinder head 110 as shown in FIGS. 7 and 8 is used, which in this case is designed for two cylinders and has an outlet chamber 112 and from the outlet chamber 112 through one Wall portion 114 separated a first inlet chamber 116 and a second inlet chamber 118, which in turn are separated by an intermediate wall 120.

The inlet chamber 116 is assigned to a cylinder 62 of the main compressor stage 66, while the inlet chamber 118 is assigned to the cylinder 62 of the additional compressor stage 70. For this reason, the inlet chamber 118 is also directly provided with a connection flange 122 for the auxiliary port 72, while the inlet chamber 116, the refrigerant is supplied via the usual, provided in the housing inlet channels.

In addition, the outlet chamber 112 is also provided with a connection flange 124 for the high-pressure connection 14.

In order to heat the refrigerant to be aspirated into the inlet chambers 116 and 118 as little as possible by means of the refrigerant flowing into the outlet chamber 112 in a refrigerant compressor according to the invention, the wall region 114 separating the outlet chamber 112 from the inlet chambers 116 and 118 is substantially larger than two Areas of the height of the cylinder head 110 formed separately walls 126 and 128, between which a free space 130 is provided, which isolates the walls 126 and 128 relative to each other and thus also the outlet chamber 112 with respect to the inlet chambers 116 and 118 thermally insulated.

The two walls 126 and 128 only essentially unite in a wall region 132 which directly adjoins a base surface 134 of the cylinder head 110.

Preferably, the check valve 76 can be arranged in the intermediate wall 120 and thus allows in a simple manner the suction of refrigerant from the inlet chamber 116, when the inlet chamber 118 of the additional compressor stage 70 via the additional port 72, no refrigerant is supplied. In a second preferred embodiment of a refrigerant compressor according to the invention, shown in Fig. 9 and Fig. 10, the intermediate wall 120 'of the cylinder head HO ^{1 is} not provided with the check valve 76, but it is a check valve 176 provided on a valve plate 140, which on a cylinder housing 142 rests and in turn the cylinder head 110 'carries.

For this purpose, an additional opening 144 is provided in the valve plate 140, which is congruent with a provided in the cylinder housing 142 and branched off from the inlet channel 148 connecting channel 174, and opens into the inlet chamber 118 for the cylinder 62 of the additional compressor stage 70.

The opening 144 can be closed by a valve tongue 178 of the check valve 176, which is arranged on a side of the valve plate 140 facing the inlet chamber 118 and is additionally secured by a catcher 180.

The inlet chamber 116 of the main compressor stage 66 is supplied via an inlet channel 148 with the low-pressure port 52 supplied refrigerant, wherein in the valve plate 140 a congruent with the inlet channel 148 arranged opening 150 is provided, via which the refrigerant from the inlet channel 148 into the inlet chamber 116th transgresses.

Thus, as shown in FIG. 10, there is a simple possibility of associating with the valve plate 140 not only inlet ports 152 of the main compressor stage 66 and inlet ports 154 of the booster stage 70 associated but not directly visible in FIG. 10 intake valves and also on the valve plate 140, the corresponding exhaust valves 156 and 158 to arrange, but in the same way, and preferably with the same structure as the exhaust valves 156 and 158, to provide the check valve 176 so that it can be mounted in a simple manner and in terms of its valve characteristics in the same Can be optimized as the exhaust valves 156 and 158.

Claims

PATENT CLAIMS

A refrigeration system comprising a refrigerant circuit (10) in which a main mass flow (56) of a refrigerant is guided, a high-pressure side, refrigerant-cooling heat exchanger (18) arranged in the refrigerant circuit (10), an expansion cooling device (24, 28) arranged in the refrigerant circuit, which in the active state cools the main mass flow (56) of the refrigerant and thereby generates an additional mass flow (86) of gaseous refrigerant, a liquefied refrigerant reservoir (30) arranged in the refrigerant circuit (10), at least one liquefied expansion unit (40) arranged in the refrigerant circuit Refrigerant of the main mass flow (56), which has an expansion element (44) and a downstream low-pressure side, cooling capacity heat exchanger (46), and at least one in the refrigerant circuit (10) arranged refrigerant compressor (12) having a main compressor stage (66 ) un d comprises at least one auxiliary compressor stage (70) driven jointly with the main compressor stage (66), which compresses both refrigerants to high-pressure PH, the main compressor stage (66) and the at least one additional compressor stage (70) being usable so that either the main compressor stage (66) Refrigerant from the main mass flow (56) and the additional compressor stage (70) refrigerant from the additional mass flow (86) or the main compressor stage (66) and the additional compressor stage (70) compress refrigerant from the main mass flow (56), characterized in that in the refrigerant circuit (10) at least two refrigerant compressors (12) are arranged, which are individually connectable for compressing the main mass flow (56) that at least two of the refrigerant compressor (12) each have at least one additional compressor stage (70) that each of Additional compressor stages (70) optionally for compressing refrigerant from the main mass flow (56) or for compressing refrigerant from the additional mass flow (86) can be used and that a controller (90) is provided, with which in a first operating mode depending on the operating conditions Such number of additional compressor stages (70) for compressing refrigerant from the additional mass flow (86) is switchable, that the Expansϊonskühleinrichtung (24, 28) liquefies the main mass flow (56) and reduces its enthalpy.

2. Refrigeration system according to claim 1, characterized in that the expansion cooling means (24, 28) reduces the enthalpy of the main mass flow (56) by at least 10%.

3. Refrigeration system according to claim 2, characterized in that the expansion cooling means (24, 28) reduces the enthalpy of the main mass flow (56) by at least 20%.

4. Refrigeration system according to one of the preceding claims, characterized in that the first operating mode corresponds to a supercritical operation.

5. Refrigeration plant according to one of the preceding claims, characterized in that the expansion cooling means (24, 28) converts the main mass flow (56) into a thermodynamic state whose pressure and enthalpy values are lower than those of a maximum (98) of the saturation curve (96). ,

6. A refrigeration system according to claim 5, characterized in that the of the expansion cooling means (24, 28) caused pressure and enthalpy of the main mass flow (56) near the saturation curve (96) in the enthalpy / pressure diagram.

7. refrigeration system according to claim 6, characterized in that the of the expansion cooling means (24, 28) caused pressure and enthalpy of the main mass flow (56) are substantially on the saturation curve (96) of the enthalpy / pressure diagram.

8. Refrigeration plant according to one of the preceding claims, characterized in that the expansion cooling means (24, 28) an expansion valve (24) for expansion of refrigerant to an intermediate pressure PZ and that the intermediate pressure PZ of the expansion cooling means (24, 28) by switching the appropriate Number of additional compressor stages (70) is adjustable.

9. refrigeration system according to claim 8, characterized in that the expansion valve (24) refrigerant of the main mass flow (56) and refrigerant of the additional mass flow (86) expands to the intermediate pressure PZ.

10. refrigeration system according to claim 8 or 9, characterized in that the expansion cooling means (24, 28) the reservoir (30) for the liquid refrigerant of the main mass flow (56) mitumfasst.

11. Refrigeration system according to one of the preceding claims, characterized in that in the second operating mode, the expansion cooling device (24, 28) is in the inactive state and does not cause cooling of the main mass flow (56).

12. Refrigeration system according to one of the preceding claims, characterized in that in the second operating mode, all additional compressor stages (70) compress refrigerant of the main mass flow (56).

13. Refrigeration plant according to one of the preceding claims, characterized in that in the second operating mode in the reservoir (30) liquid refrigerant of the main mass flow (56) is at high pressure PH.

14. Refrigeration system according to one of claims 11 to 13, characterized in that the second operating mode corresponds to a subcritical operation.

15. Refrigeration plant according to one of the preceding claims, characterized in that the controller (90) controls the refrigerant compressor (12) according to the required cooling capacity (48).

16. Refrigeration system according to one of the preceding claims, characterized in that with the controller (90) corresponding to the required cooling capacity (48), the refrigerant compressor (12) individually switched on or off.

17. Refrigeration system according to one of the preceding claims, characterized in that each refrigerant compressor (12) with additional compressor stage (70) is dimensioned such that the compressed by the additional compressor stage (70) mass flow of refrigerant of the additional mass flow (86) maximum of the main compressor stage ( 66) corresponds to a compressed mass flow of refrigerant of the main mass flow (56) in this refrigerant compressor (12).

18. Refrigeration system according to one of the preceding claims, characterized in that the refrigerant compressor (12) with additional compressor stage (70) are dimensioned so that the additional compressor stages (70) of different refrigerant compressor (12) compress different mass flows of refrigerant of the additional mass flow (86).

19. Refrigeration system according to one of the preceding claims, characterized in that the refrigerant compressor (12) with additional compressor stage (70) are reciprocating compressors.

20. Refrigeration system according to claim 19, characterized in that each of the refrigerant compressor (12) with additional compressor stage (70) has at least one cylinder (62) for the additional compressor stage (70) and at least one cylinder (62) for the main compressor stage (66).

21. Refrigerant compressor according to claim 19 or 20, characterized in that in each refrigerant compressor (12) with additional compressor stage (70), the number of cylinders (62) for the main compressor stage (66) is greater than the number of cylinders (62) for the additional compressor stage (70).

22. Refrigeration system according to the preamble of claim 1 or any one of claims 19 to 21, characterized in that of the refrigerant compressors (12) with additional compressor stage (70) the additional compressor stages (70) of different refrigerant compressor (12) have a different stroke volume.

23. A refrigeration system according to claim 22, characterized in that in each refrigerant compressor (12) with additional compressor stage (70) the ratio of the stroke volume of the additional compressor stage (70) to the displacement of the main compressor stage (66) against at least one of the other refrigerant compressor (12) with additional compressor stage (70). 70) is different.

24. Refrigeration plant according to one of the preceding claims, characterized in that in the first operating mode the reservoir (30) for liquefied refrigerant operates at an intermediate pressure PZ and between the high-pressure side, the refrigerant cooling heat exchanger (18) and the reservoir (30) for liquefied Refrigerant an additional expansion unit (100) with an expansion device (104) under a downstream cooling capacity (108) emitting heat exchanger (106) is provided.

25. Refrigeration system according to one of claims 19 to 24, characterized in that the refrigerant compressor (12) cylinder heads (110), in which outlet chambers (112) and inlet chambers (116, 118) are formed substantially thermally decoupled.

26. Refrigerant according to one of the preceding claims, characterized in that a check valve (76, 176) for connecting an inlet chamber (118) of the additional compressor stage (70) with the low-pressure connection (52) of the main compressor stage (66) is provided.

27. Refrigeration system according to claim 26, characterized in that the check valve (176) is provided in a valve plate (140) of the respective refrigerant compressor (12).

28. Refrigeration system according to claim 27, characterized in that in a cylinder housing (142) of the connecting channel (174) between the low pressure port (52) and the check valve (176) extends.