The present invention relates to the field of refrigeration technology and air-conditioning technology, including heat pump systems, based on a thermodynamic cyclic process. It refers to a system for refrigeration, heating or air-conditioning technology according to the preamble of claim 1 and to a method for operating such a system.
On the one hand, it is known, in refrigeration technology, to have dry expansion operation, in which the working medium or refrigerant experiences a pressure reduction via an injection valve and changes from the liquid state into a liquid/vapor mixture, and in the following evaporator evaporates completely, in order then to leave the evaporator in the form of slightly superheated vapor, and thus, by heat absorption, cools down a second medium (for example, a brine).
On the other hand, thermosiphon operation is known, in which the refrigerant is delivered as liquid from an equalizing or separating vessel to the evaporator either by means of gravity or with the aid of a pump. Upon exit from the evaporator, it is perfectly possible that the vapor still contains liquid fractions, so therefore, as a rule, no superheating of the refrigerant occurs at the evaporator exit.
Furthermore, U.S. Pat. No. 5,243,837 discloses a refrigeration system (see FIG. 1 there), the basic set-up of which is reproduced in FIG. 1. The known refrigeration system 10 of FIG. 1 comprises a refrigerant circuit, in which a compressor 12, a condenser 11, a heat exchanger 13, an expansion or injection valve 15 and an evaporator 14 are arranged one behind the other in the direction of flow. In the condenser 11, the compressed refrigerant is liquefied by heat exchange with a medium supplied and discharged via the connecting lines 16 and 17. Air which flows through the condenser 11 may also be considered as a medium. In the heat exchanger 13, the refrigerant liquefied in the condenser 11 is supercooled by heat exchange with the suction vapor flowing to the compressor 12, while the suction vapor is itself superheated. The liquid supercooled refrigerant which is under pressure is expanded in a controlled manner in the expansion valve 15, the volume flow of the refrigerant being regulated. The expanded liquid/vapor mixture evaporates in the evaporator 14 and at the same time cools a secondary medium supplied and discharged by means of connecting lines 18, 19. However, the cold of evaporation may also be transferred directly via a cold surface for cooling a space. The evaporated refrigerant leaves the evaporator in slightly superheated form. The (internal) supercooling of the liquid refrigerant prior to expansion in an expansion valve 15 increases the efficiency of the refrigeration system. However, the associated subsequent high superheating of the evaporated refrigerant in the heat exchanger 13 has an adverse effect on the efficiency of the compression process.
In U.S. Pat. No. 5,243,837 mentioned, therefore, it was proposed (see FIG. 2 there) to carry out a further supercooling of the liquid refrigerant directly in the evaporator which has an integrated supercooler. Such a solution is reproduced in FIG. 2 where the evaporator 21 is additionally equipped with an internal supercooler in the refrigeration system 20.
To increase the profitability of refrigeration systems of this type, it was proposed by the applicant in WO-A1-2004/020918 to use refrigeration systems of modular construction with compressors of variable rotational speed, in which, depending on the refrigerating capacity required, individual modules are cut in or cut out or are varied individually in terms of their power, in order to compensate the power jumps caused by entire modules being cut in and cut out. The modular construction of the refrigeration system results, for each individual module, in especially beneficial small refrigerant fillings. Furthermore, the alternating use of a plurality of modules achieves a redundancy which assists in avoiding interruptions in the processes relying upon the generation of cold. The individual modules in this case have the set-up, shown in FIG. 3, of a refrigeration system 30 with a two-stage evaporator 22 which comprises a first evaporator stage 23 and a following second evaporator stage 24 in the form of an internal heat exchanger (IHE).
Under practical conditions, all these systems have more or less serious disadvantages: dry expansion systems have the advantage of a simple type of construction and low refrigerant contents. The system efficiency is influenced essentially by as low an evaporation superheating as possible and by as high an evaporation temperature as possible. This, however, is a disadvantage for the compressor which requires a correspondingly high superheating (charging efficiency improvement, lubrication, etc.). The point of intersection of these two contrary requirements (low superheating to the evaporator, high superheating for the compressor) gives the optimal system characteristic curve (the most efficient operation).
It was proposed by the applicant, then, in WO-A1-2005/073645, to break through this dependence between the lowest possible superheating for the evaporator and high superheating for the compressor. An attempt is made, in this context, to operate the process for a given refrigerating capacity Q0 by means of the lowest physically possible mass flow required for this purpose, thus leading to considerable economic and energy-related benefits. The solution proposed there, may be employed both in dry expansion systems with a following internal heat exchanger (according to FIG. 1), that is to say with heat exchange (13) between the refrigerant liquid line upstream of the expansion valve 15, on the one hand, and the suction vapor downstream of the evaporator 14, on the other hand, and in two-stage evaporation systems according to FIG. 3.
However, what is typical of all conventional systems, depending on the operating conditions, is that there are relatively high temperature fluctuations upstream of the expansion valve (injection valve) 15 and upstream of the compressor 12 on the refrigerant side. These temperatures of the refrigerant (upstream of the injection valve 15 and upstream of the compressor 12) have hitherto not been kept constant or regulated exactly. Often, if at all, only the high or suction pressure has been regulated and/or kept constant. This has led to greater or lesser fluctuations and feedbacks (oscillation) of the refrigeration system and therefore to unstable control loops and loss of efficiency. The main factors for these fluctuations, on the one hand, are the inlet vapor content into the evaporator, which varies with the varied temperature of the refrigerant and which has effects on the injection-valve performance and evaporator performance and the control behavior of the injection valve and its performance, or the conveyed refrigerant mass flow. On the other hand, actions in the case of the suction vapor also arise at the inlet into the compressor 12, where the varied temperature has an influence on the conveying volume of the compressor 12, that is to say, again, a conveyed mass flow, on account of the specific volume assigned to the respective temperature (and to the respective pressure). These mass flows which vary constantly as a result of the temperature changes introduce greater or lesser disturbance factors into the control loop of the refrigeration system, thus leading to fluctuations in the process and therefore to performance reductions.
A stable operation of the system is achieved in WO-A1-2005/073645 in that:
- the temperature of the refrigerant upstream of the injection valve 15 is kept constant at a defined temperature value; or
- the temperature of the refrigerant upstream of the compressor is kept constant at a defined temperature value;
- these two measures being used each alone or in combination with one another.
- The three first measures with dry expansion valve control are operated conventionally according to MSS (most minimal stable signal) with or without IHE (internal heat exchanger).
- The injection valve 12 is regulated by means of the temperature between the liquid line upstream of the injection valve 12 and pressure measurement downstream of the injection valve 12, what is known as two-stage evaporator regulation (according to WO-A1-2004/020918).
These measures, such as keeping the temperature of the refrigerant liquid constant upstream of the injection valve 12, keeping the suction vapor temperature constant upstream of the compressor 12 and the two-stage evaporator process (with corresponding regulation of the injection valve 12) lead, alone or in any desired combination, to stable operation of the refrigeration systems (also in the event of pronounced performance changes). If, in this case, according to FIG. 3 a two-stage evaporator 22 is used, the lowest possible temperature differences between the medium to be cooled, on the one hand, and the evaporation temperature To (at suction pressure), on the other hand, can additionally be achieved. This temperature difference may in any event be lower than when the refrigerant leaves the evaporator 14 in “superheat” form in the case of dry expansion operation.
As a result of the more stable operation which is acquired, energy and cost savings are obtained, and it becomes possible, especially in combination with the two-stage evaporation technology (FIG. 3), to operate processes with substantially lower temperature differences of the media to be cooled in relation to the respective evaporation temperatures. As a result, processes which are not possible in this way today can be operated in a simple and cost-effective way.
The known solution for stabilizing the refrigeration system allows a marked improvement in the operating behavior and in efficiency, at the same time also places requirements upon the regulation of the system.
The object on which the present invention is based, therefore, is to provide a system for refrigeration, heating or air-conditioning technology, particularly a refrigeration system, in which stable operating conditions can be achieved in an especially simple way, and also to specify a method for operating it.
The object is achieved by means of the whole of the features of claims 1 and 8. An essential point of the invention is that, to keep the temperature of the liquid working medium or refrigerant constant upstream of the expansion valve, means are provided which couple the liquid working medium flowing to the expansion valve thermally to the working medium flowing from the expansion valve to the evaporator.
A preferred refinement of the invention is distinguished in that the means for keeping the temperature of the liquid working medium constant upstream of the expansion valve comprise a stabilizer in the form of a heat exchanger, through which the working medium flowing from the expansion valve to the evaporator flows on one side and through which the liquid working medium flowing to the expansion valve flows on the other side. In particular, the working medium flows through the stabilizer in cocurrent or countercurrent. However, other types of routing of the streams of the working medium in the stabilizer may likewise be envisaged.
According to another refinement of the invention, the evaporator is followed by an internal heat exchanger, in which, on one side, the working medium coming from the evaporator is re-evaporated and/or superheated and, on the other side, the working medium coming from the condenser is supercooled before entry into the stabilizer. In particular, the internal heat exchanger is designed as a thermally long heat exchanger.
The performance of the system can be increased in that an external supercooler is inserted between the condenser and the internal heat exchanger and/or in that a waste heat utilization exchanger is arranged between the compressor and the condenser. The heat energy obtained in the waste heat utilization exchanger is in this case, as a rule, beneficial to a second process (for example, service water, heating).
According to a preferred refinement of the method according to the invention, the internal heat exchanger is selectively operated, as a function of the inlet temperature of the liquid working medium into the internal heat exchanger, solely as a superheater for the working medium flowing to the compressor or as a further evaporator stage. It is especially efficient if in this case, in the system, the highest possible power in the evaporator is transmitted to a secondary medium by means of the lowest possible mass flow on the cold side.
The invention will be explained in more detail below by means of exemplary embodiments, in conjunction with the drawing in which:
FIG. 1 shows a refrigeration system according to the prior art for dry expansion operation with subsequent superheating/supercooling;
FIG. 2 shows a refrigeration system based on FIG. 1 and known from the prior art, with additional supercooling integrated in the evaporator;
FIG. 3 shows a refrigeration system according to the prior art with a two-stage evaporator;
FIG. 4 shows a refrigeration system according to a first exemplary embodiment of the invention with a stabilizer arranged directly on the expansion valve;
FIG. 5 shows a refrigeration system, based on FIG. 4, according to a second exemplary embodiment of the invention, with an additional waste heat utilization exchanger and with an external supercooler;
FIG. 6 shows a pressure/enthalpy graph of a cyclic process which is operated by means of a system according to FIG. 5 and in which the internal heat exchanger (26) works as a straightforward superheater; and
FIG. 7 shows a pressure/enthalpy graph of a cyclic process which is operated by means of a system according to FIG. 5 and in which the internal heat exchanger (26) works as a third evaporator stage.
FIG. 4 reproduces a refrigeration system according to the first exemplary embodiment of the invention in a greatly simplified diagram. The refrigeration system 40 has a working-medium circuit or refrigerant circuit, in which a compressor 12, a condenser 11, an expansion valve 15 and an evaporator 14 are arranged one behind the other in the direction of flow of the working medium or refrigerant. The refrigerant (for example of the type R134a) is compressed in the usual way in the compressor 12, is then liquefied in the condenser by heat exchange with an external medium (air, water or the like) and is then routed to the (usually controllable) expansion valve 15 where it is expanded in a controlled manner. The expanded liquid refrigerant, which may already have vapor fractions here, is supplied to the evaporator 14, where, by evaporation, it absorbs heat from a secondary medium supplied and discharged via connecting lines 18, 19 or cools this medium.
Since, after the expansion of the refrigerant in the expansion valve 15, the relatively constant evaporation temperature of the refrigerant is established at the pressure which prevails there, this temperature can be used in order to stabilize the temperature of the refrigerant upstream of the expansion valve 15. For this purpose, according to FIG. 4, a stabilizer 25 in the form of a heat exchanger is inserted between the expansion valve 15 and the evaporator and couples the refrigerant stream to the expansion valve 15 thermally to the refrigerant stream routed from the expansion valve 15 to the evaporator 14. As a result of this stabilizing coupling, the control fluctuations which otherwise occur in the system can be largely avoided in a simple way.
The operating behavior of the refrigeration system 40 is especially beneficial if the evaporator 14 is followed by an internal heat exchanger (IHE) 26 which can work as a second or third evaporation stage (FIG. 7) or as a straightforward superheater (FIG. 6). The refrigerant (vapor or liquid/vapor mixture with a low liquid fraction) coming from the evaporator 14 is fed on one side, through the internal heat exchanger 26, to the compressor 12. On the other side of the internal heat exchanger 26, condensed refrigerant flows to the stabilizer 25 and is in this case supercooled in the internal heat exchanger 26.
In the pressure (p)/enthalpy (h) graph of FIG. 6, in which the phase limit curve of an exemplary refrigerant and (by dashes) a typical curve of constant temperature are depicted, the following steps arise for the associated cyclic process A-B-C-D: in process step A, the superheated refrigerant emerging from the internal heat exchanger 26 is compressed. In process step B, the compressed refrigerant is deheated, condensed, supercooled externally and internally and lastly further lowered in temperature in the stabilizer 25. In process step C, the liquid refrigerant is expanded. In the last process step D, the expanded refrigerant is partially evaporated in the stabilizer 25, is completely evaporated and slightly superheated in the evaporator 14 and is further superheated in the internal heat exchanger 26, in order then to arrive at the compressor again.
In the comparable pressure/enthalpy graph of FIG. 7, by contrast, the internal heat exchanger 26 acts as a third evaporator stage with correspondingly lower superheating, thus leading to a displacement of process steps A′ and B′.
The stabilizer 25 is as it were the first evaporator stage and is always in operation and, depending on the exchanger quality (“thermal length”), cools the refrigerant liquid, in practice, down to the evaporation temperature (the “thermal length” of the heat exchanger is in this case a measure of the approximation of the initial temperatures on the primary or secondary side of the heat exchanger to the respective inlet temperatures of the (opposite) secondary or primary side; in a heat exchanger with a long thermal length, these two temperatures are approximately equal). The second evaporator stage is formed by the evaporator 14 itself. A third evaporator stage is obtained when the IHE 26 is used for the evaporation of residual liquid, as shown in the graph of FIG. 7.
By means of the stabilizer 25, downstream of the expansion valve 15, the refrigerant stream has virtually no or only an insignificant vapor fraction when it enters the first evaporator stage (stabilizer 25). This fact also affords benefits with regard to incorrect distribution in the case of, for example, plate heat exchangers.
Depending on the inlet temperature of the refrigerant liquid into this first evaporator stage (depending on the preceding process or on the inlet temperature into the IHE or the third evaporator stage 26, or the quality of the heat exchanger, etc.), more or less liquid is evaporated for stabilization. This evaporation process does not entail any direct increase in performance of the system and serves merely for keeping constant the process (indirect performance increase) which otherwise, due to the use of a thermally long heat exchanger as IHE, becomes highly unstable and scarcely controllable, especially when evaporation takes place in this stage (third evaporation stage; FIG. 7).
Of course, this type of stabilization is not restricted to systems with two-stage evaporation (with the IHE 26 as the second evaporation stage), but also offers advantages in all conventional evaporation processes, such as, for example, dry expansion.
The second evaporation stage, the actual evaporator 14 which cools down a secondary medium (water, brine, air, etc.), is thus operated with a fraction, varying according to the process, of already evaporated refrigerant (approximately 0-45%, depending on the refrigerant). The outlet conditions of the refrigerant out of this second stage may be different, depending on the refrigerant liquid inlet temperature into the IHE 26: the emerging refrigerant may be superheated in gaseous form (FIG. 6) or may be in the form of a wet vapor (FIG. 7).
The aim is to leave this second evaporator stage 14 with minimal superheating (1-8 K) if the least temperature differences of the secondary medium (brine, etc.) are to be ignored and therefore the evaporation temperature is not also lowered (otherwise, it will have emerged from the evaporator 14 on the boundary line with approximately saturated vapor without superheating). The third stage, the IHE 26, then serves solely for superheating the working medium or refrigerant (FIG. 6).
If, because of the conditions of use mentioned above, the compressor does not achieve this or the hot-gas temperatures are too high, the refrigerant must leave the second evaporator stage (evaporator 14) with a fraction of liquid which evaporates in the third stage and thus limits the suction gas inlet temperature of the refrigerant into the compressor 12 to a permissible value (FIG. 7). The internal heat exchanger 26 then forms, together with the evaporator 14, a two-stage evaporator (TSE). Together with the stabilizer 25 as a further evaporator stage, overall, a three-stage evaporator 25, 14, 26 is obtained. The fraction of liquid refrigerant in the last evaporator stage, the IHE 26, is in any event a loss, since this fraction of the evaporator power is not for the benefit of the secondary medium (brine, etc.) to be cooled.
The operation of a system with a stabilizer 25 and with a purely superheating internal heat exchanger IHE or third evaporation stage 26 can be described as follows: in the first evaporator stage (stabilizer 25), the refrigerant is evaporated, in order to cool the refrigerant liquid down to near the evaporation temperature and thus to obtain stable operation. In the second evaporator stage (evaporator 14), refrigerant is then evaporated, in order to transmit the highest possible power by means of the lowest possible mass flow (defined by a process in which the same mass flow flows through all the lines). In this case, a thermally long internal heat exchanger 26 is used, in order to cool the refrigerant liquid as low as possible by means of the cold suction gases.
If in this case the suction gases are heated inadmissible highly, this heating is limited by a residual evaporation of the refrigerant on the suction side in the heat exchanger 26 (FIG. 7), thus having an influence on the refrigerant liquid temperature which is lowered to a greater extent than without this residual evaporation, but the suction gas temperature is also limited and signifies a loss of refrigerating capacity in the process, as compared with straightforward suction vapor superheating without limitation of the suction gas temperature. Since stabilization already takes place as a result of the stabilizer 25, the system can be operated in both operating modes (FIG. 6 or FIG. 7): depending on requirements, the heat exchanger 26 can be operated as a “dry” superheater or as an additional third evaporation stage. Whether the first or the second operating mode prevails is determined only by the refrigerant liquid inlet temperatures into the heat exchanger 26 and the suction gas outlet temperature (limits of use of the compressors, oil, hot-gas temperature) and may therefore change constantly, for example, during the day and/or annually and also in the event of an operating point changeover (for example, the demand for brine, on the one hand, at 0° C. and, on the other hand, at −25° C.).
If, in addition, the system is constructed from a plurality of individual largely identical modules, a temperature regime of heat-pump, air-conditioning to freezing temperatures can be operated with the same module (compressor with high-power motor, electronic injection valve, correct determination of the heat exchangers), and virtually any performance spectrum can be covered by frequency control and the number of existing or cut-in modules, this being achieved by means of a technology which is always the same (identical modules). Especially high efficiency can in this case be achieved in that, in each of the modules, the highest possible power in the evaporator 14 is transmitted to the secondary medium by means of the lowest possible mass flow on the cold side. Of course, the refrigeration system may be extended by means of additional components, as is illustrated by way of example for the refrigeration system 40′ in FIG. 5. In this exemplary embodiment, a waste heat utilization exchanger 27 is additionally inserted between the compressor 12 and the condenser 11 and discharges part of the compression heat by heat exchange with an external medium supplied and discharged via the connecting lines 29 and 31. An external supercooler 28 may likewise be inserted between the condenser 11 and the internal heat exchanger 26 and causes a first supercooling of the condensed refrigerant by heat exchange with an external medium supplied and discharged via the connecting lines 32 and 33.
This multistage supercooling affords the possibility, by an external limitation of the suction gas temperatures, of limiting the suction gas temperature as a safety device. Any further cooling of the refrigerant liquid downstream of the condenser 11 and upstream of the IHE 26 by means of one or more external supercoolers 28 is in any case a power gain and should, if possible, be provided. The lower limits are determined again by the limits of use of the compressors 12. In extreme situations, too low and overall superheating could occur, and the compressor 12 could be destroyed as a result of liquid surges. Furthermore, the aim is also always to have the lowest possible condensation temperatures.
If operation in summer is assumed, the highest refrigerating capacity is often required on the hottest days. In the case of air-cooled condensers, the outside temperature determines the liquefaction pressure (temperature) and the refrigerant liquid temperature in the IHE 26. In this case, the suction gas temperatures may reach too high a value (limitation by the external supercooler 28 or evaporation in the IHE 26 as a third evaporation stage; at these high suction gas temperatures, an external supercooler 28 is the most efficient).
If the air temperature falls during the night or in winter, the system can or should be operated with lower condensation temperatures, this resulting automatically in lower refrigerant liquid temperatures in the IHE 26, so that the suction gas temperature no longer has to be limited and the IHE 26 functions as a heat exchanger without evaporation. Here, too, the external supercooler 28 can bring about an increase in power, but never more to the same extent as in summer. In this case, a different amount of refrigerant evaporates in the stabilizer 25, with or without an external supercooler 28, depending on summer or winter and day or night operation.
It would be appreciated that the explanations given above referred only to a refrigeration system, but that the solution principles illustrated can readily be applied, within the scope of the invention, to other thermodynamic systems of air-conditioning and heating technology. The types of construction of the heat exchangers used may in this case be as desired (for example, plate heat exchangers or the like).
Overall, by virtue of the invention, a highly efficient thermodynamic system with an especially stable operating behavior, which is distinguished by a very simple construction, is obtained.
LIST OF REFERENCE SYMBOLS
- 10, 20, 30, 40, 40′ Refrigeration system
- 11 Condenser
- 12 Compressor
- 13 Heat exchanger
- 14 Evaporator
- 15 Expansion valve (injection valve)
- 16, 17 Connecting line
- 18, 19 Connecting line
- 21 Evaporator
- 22 Two-stage evaporator
- 23 First evaporator stage
- 24 Second evaporator stage
- 25 Stabilizer
- 26 Internal heat exchanger (IHE)
- 27 Waste heat utilization exchanger
- 28 External supercooler
- 29, 31 Connecting line
- 32, 33 Connecting line