EP0330198B1 - Heat exchanger as an injection evaporator for a refrigeration machine - Google Patents

Abstract

A heat exchanger (1) provided as an injection evaporator for a refrigeration machine comprises a first section (11) with a broad flow cross-section, in which the refrigerant mixture can spread, at the bottom in the admission region, under the influence of gravity and subsequently rises in the manner of a flooded evaporator. A second section (12) follows, in which the flow cross-section is so narrow that a gas speed sufficient for oil transport is achieved. The boundary between the two sections lies below the region of complete evaporation of the refrigerant so that the oil reaches the second section mixed with liquid refrigerant. <IMAGE>

Description

In the case of evaporators for refrigeration machines, which also include heat pumps in the present context, a distinction is made between those for flooded operation and those for injection operation. In the case of flooded evaporators (DE-A-3309979), the refrigerant is supplied in the liquid phase and discharged with mixed phases, a separator being required for phase separation. Since the pressure loss and the static pressure difference within the evaporator are small, the evaporation temperature, which is determined by the pressure of the refrigerant, is essentially constant over the height of the evaporator. This is favorable for the energy balance and for those applications in which the controllability of the temperature profile in the medium to be cooled is important, for example in the ice-free cooling of water near freezing point with a high heat flux density. In the case of smaller systems in particular, the construction and control expenditure for the separator and the need for a special oil return from the separator to the compressor are disadvantageous. The latter results from the fact that the oil required for lubricating the compressor is partly removed with the refrigerant, in the The condenser in the liquid portion of the refrigerant dissolves or mixes with it so that it passes from the evaporator into the separator and accumulates there as a result of distillation in the liquid portion. The plate heat exchanger for flooded evaporators (DE-A 3147378, US-A 2028213) are characterized by a refrigerant space in which the evaporating refrigerant mixture rises against gravity and which have such a large flow cross section that the mixture spreads across the width of the heat exchanger can distribute sufficiently evenly due to the effects of gravity.

In evaporators with injection operation (FR-A-2549585), the refrigerant is supplied to the evaporator without prior separation as a mixture of the liquid and the gaseous phase and completely evaporated, whereby to reliably avoid liquid hammer in the compressor and to control the injection valve designed as a thermostatic expansion valve a certain amount of overheating has to be accepted. The expenditure on equipment for a refrigeration machine in injection mode is lower, which is why this design is often preferred for smaller systems (for example below 200 KW). However, it has the disadvantage that there is a considerable pressure drop on the evaporator as a result of a forced flow and thus a considerable temperature difference on the refrigerant side, which is not only economically undesirable, but also for use with precise temperature control of the medium to be cooled, for example for heat pumps or ice water systems with near freezing point lying water temperature. The forced flow is intended, on the one hand, to ensure that the compressor lubricating oil, which remains after evaporation of the refrigerant as the only liquid component in the evaporator, is discharged through a sufficient gas velocity. On the other hand, it is intended to ensure that the heat exchanger surfaces are wetted evenly despite the small amount of liquid (only a few% by volume at the evaporator inlet) are. With comparatively low heat flow densities, such as are present in heat exchangers with a gaseous external medium, the latter condition can be met with moderate speeds and a moderate pressure drop (FR-A 2549585). However, high speeds with a correspondingly high pressure loss are required if high heat flow densities are to be achieved, as is possible with a liquid external medium (for example with trickle coolers). It is customary (GB-A 1286446; DE-A 3536325) to increase the flow cross section in accordance with the increase in the flowing mixture volume in the course of the heat exchanger section. A contrary, very old proposal for an arrangement with falling refrigerant flow (DE-C 161027) contradicts decades of practice and relevant physical considerations.

It is known in the case of injection evaporators to provide a plurality of separate sections with different flow cross sections within a plate evaporator; However, these only form a structural unit, but not a functional unit, because they are each separately provided with an injection valve, which requires a high level of control engineering and construction. The basic disadvantage that a high pressure drop and thus a large temperature difference occurs in injection evaporators cannot be avoided by these means.

In the case of a heat exchanger which, as an injection evaporator for a refrigeration machine, enables a high heat flow density and the removal of oil which has been removed, the invention aims to achieve a low pressure drop, comparable to that of a flooded evaporator.

The solution according to the invention consists in a method for operating a refrigeration machine, which is characterized in that the refrigerant passes from a first section of the heat exchanger with a second flow cross section and gravity distribution of the rising refrigerant into a second section with a narrower flow cross section in the not yet fully evaporated state, wherein the refrigerant in the second section has a speed sufficient to convey failed oil.

In the first section of the heat exchanger there is therefore an operating behavior similar to that of a flooded evaporator. The pressure drop in this section is therefore very low. This enables precise temperature control. When cooling water, ice-free operation down to, for example, 0.5 ° C is therefore possible. The heat economy is improved. A higher pressure and temperature drop occurs only within the second evaporator section.

The invention is based on the knowledge that, in the case of an injection evaporator, a significantly increased speed of the refrigerant is only required in the area in which the complete evaporation of the liquid phase of the refrigerant takes place, so that there is uniform overheating without entrained residues of the liquid phase and because also it is only here that the concentration of the oil in the liquid refrigerant component becomes so great as a result of refrigerant evaporation that there is a risk of oil failure. Furthermore, the invention is based on the knowledge that even in the injection mode, the liquid portion in the evaporator is still large enough to ensure adequate liquid wetting of the evaporator inner surfaces even at a relatively low medium speed. Even if the proportion of liquid entering the evaporator is only of the order of 10% by volume, it still comprises around 80% by weight. This is sufficient for a sufficient distribution of gravity of the incoming mixture at least in the lower insertion area of the first section. A completely even density distribution is not necessary because the wetting also through the subsequent gas development and the resulting turbulence is promoted.

An evaporator for a household refrigerator is known (US Pat. No. 2,414,952) which, as the refrigerant flow increases, comprises a first section with a number of channels connected in parallel and a second, higher section which is formed by only one channel. The only information about the flow cross-sections is that they should be so large in the second section that liquid refrigerant can flow back to the first section against the gas flow. The gas velocity is therefore not high enough to carry away any oil that falls out. This then accumulates in the evaporator.

The size of the first section in relation to the second depends on the type of refrigerant and the type and amount of compressor lubricant to be expected. The more miscible the lubricant is even with small amounts of the liquid refrigerant, the safer it can be that the lubricant is diluted so much by the liquid refrigerant in the end region of the first evaporator section and the viscosity is therefore reduced so much that it is transported with sufficient security. The temperature also plays a role. A calculation has shown that when using the refrigerant Frigen R 22 using oil as a lubricant for a piston compressor, the oil is transported in the evaporator's liquid phase with sufficient certainty as long as the liquid component is not less than about 20 Weight percent of the refrigerant. This in turn means that when the evaporator is designed as a vertical plate evaporator, no more than the upper third to quarter need to be designed in the form of narrower, meandered flow channels. The first section takes accordingly about two thirds to three quarters of the evaporator height.

The first section can be designed in the manner of a flooded evaporator, namely as an essentially uniform space of comparatively large horizontal cross-section, in which the mixture can flow essentially vertically upwards. Constrictions can be limited to the purpose of equalizing the flow movement over the entire cross-section and improving the heat transfer, namely preferably in the form of welded connections between the plates delimiting the flow space, which are alternately offset with respect to the vertical direction and as short welding distances, welding spots or the like can be trained. Each welding point forms an evaporation core.

The evaporator can be used in immersed mode, i.e. in the container filled with the liquid to be cooled. However, its advantages are particularly evident when it is designed as a sprinkler evaporator. In this case, as is known per se, a collector forming the upper end of the plate can be thickened horizontally transversely to the plane of the plate in order to give the falling lines of the water flowing down from the outside a stronger horizontal component, which causes the water to spread out as a uniform film. If, in the context of the invention, the second evaporator section is designed in the form of horizontal channels which are alternately connected at both ends, it can be provided that the uppermost of these channels is horizontally thicker than the following channels in order to perform this function.

Particular attention is paid to the overheating of the refrigerant gas and the extraction of the oil in these channels when the refrigerant has completely or predominantly evaporated. In order to the oil, which mainly collects in the lower area of the horizontal channels, cannot flow back against the gas flow, according to the invention an increased threshold is provided on the inlet side at the lower channel boundary. At the rear end of each horizontal channel in the direction of flow, devices can be provided which facilitate the entrainment of the oil to the next upper channel, for example a narrowing of the flow cross section to increase the gas velocity and to intensify the delivery effect. The connecting piece discharging the refrigerant gas from the evaporator is expediently connected near the lower limit of the associated channel of the second section, so that the oil does not have to be raised again.

The injector supplying the evaporator is expediently a thermostatic control valve which is designed with a connection to an overheated refrigerant gas leading from the evaporator so that the refrigerant supply to the evaporator is regulated as a function of the superheating temperature. This ensures that the thermally undesired overheating area of the evaporator remains as small as possible.

The refrigeration machine is set so that the refrigerant essentially has a sufficient liquid portion to prevent lubricant failure when the boundary between the first and the second evaporator section is reached. Non-compliance with this condition is permitted for a short time, namely for such short periods of time that the lubricant cannot accumulate excessively in the first evaporator section and thus endanger the lubrication of the compressor.

The goal is achieved in this way, the total pressure loss and the Reduce evaporation temperature change to about a third. For example, while in a conventional injection evaporator plate for a trickle cooler with a heat flow density of 3 KW / m², the change in the evaporation temperature in a typical application is no longer acceptable at approx. 9 ° C, thanks to the invention it drops to approx. 3 ° C, whereby the greatest drop in temperature is reduced to a small, upper section of the evaporator, in which the water temperature is still comparatively high and the risk of ice formation is therefore low. This makes it possible for the first time to use an injection evaporator to cool water near freezing. In the lower, first evaporator section, the offset weld seam arrangement achieves a better distribution of the water film, an increased heat transfer due to a higher degree of turbulence and thereby a higher wall temperature, which likewise improves the possibilities of cooling closer to the freezing point without ice build-up.

In ice storage mode (both in the sprinkling process and under water), the controlled and more uniform temperature curve in the evaporator plate guarantees a more uniform growth of the ice than is possible with injection evaporators of conventional design. Ice blasting from the plates by means of hot gas injection can also be used.

Fig. 1

eine erste Ausführungsform des Verdampfers im Querschnitt mit einer schematischen Darstellung der Kältemaschine,

Fig. 2

Einzelheiten des Verdampferaufbaus im zweiten Abschnitt,

Fig. 3

eine zweite Verdampferausführung und

Fig. 4

den Temperaturverlauf über die Höhe des Verdampfers im Vergleich mit anderen Verdampferbauarten.

Further details of the invention are shown below with reference to the drawing. It shows:

Fig. 1

a first embodiment of the evaporator in cross section with a schematic representation of the refrigerator,

Fig. 2

Details of the evaporator structure in the second section,

Fig. 3

a second evaporator version and

Fig. 4

the temperature curve over the height of the evaporator in comparison with other evaporator types.

1, the refrigerator consists of evaporator 1, compressor 2, condenser 3 and thermostatic expansion valve 4, the pulse line 5 of which connects to a temperature sensor 6, which is arranged on line 7, which supplies the superheated gas from evaporator 1 to compressor 2.

The evaporator 1 is a vertical plate evaporator, through which the refrigerant flows from bottom to top. It consists of a first section 11 and a second section 12. The first section of the refrigerant flows substantially uniformly over its entire width from bottom to top similar to a flooded evaporator, with horizontal welding sections 13 arranged offset to the vertical direction ensuring a uniform flow and good heat transfer . Since the available flow cross-section is large, the pressure loss is low.

In the second section 12, the flow path is formed by a meandering channel 14, which is composed of a plurality of horizontal channel sections which are alternately connected at the ends and which are formed by horizontal weld seams 15 connecting the sheets forming the plate evaporator. The cross section of the channel 14 is significantly smaller than that of the first evaporator section. The flow cross section in the first section is preferably at least three times, better at least five times and usually at least ten times larger than in the second section, which results in a correspondingly higher gas velocity for the second section. The evaporator is operated so that the refrigerant with a bottom Weight fraction of the liquid phase of, for example, 70% is supplied. The quantity supplied is determined as a function of the temperature of the superheated gas in line 7 by the injection valve. This ensures that the refrigerant always reaches the beginning of the second section 12 with such a large liquid fraction that the transport of the oil into the second section is ensured, where the refrigerant evaporates completely and the gas velocity is so high that the oil is entrained becomes.

So that the oil collecting in the lower region of the horizontal sections forming the channel 14 cannot flow back, a threshold 16 can expediently be provided at the beginning of the channel. Instead, it would also be conceivable to arrange the horizontal channels to fall slightly. Furthermore, baffles (not shown) can be provided in the vertical channel connections in order to intensify the gas flow there and to improve the oil transport. On the uppermost channel 17, the discharge pipe 18 is arranged near the lower boundary of the channel 17 in order to be able to discharge the oil there more easily. Furthermore, the uppermost channel 17 can be bulged more than the ones below it, in order to improve the liquid film formation on the outside of the evaporator when it is sprinkled, as is indicated by dashed lines in FIG. 2.

In the first evaporator section, the ascending movement of the refrigerant ensures that the inner surfaces are evenly wetted despite the relatively slow flow without separation phenomena. For the second evaporator section, on the other hand, the predominantly horizontal flow is considered to be advantageous, so that in those areas in which oil must be expected to separate depending on the gas velocity, this can collect in the lower area of the horizontal channels, to a lesser extent by wetting the other inner surfaces deteriorate the heat transfer. Unlike in Fig. 1 In the second section to improve oil production (in particular for liquid cooling), a falling connection of the horizontal channels can also be provided, the uppermost channel 17 being connected directly to the first section 11 by a vertical channel 20, as shown in FIG. 3. This gives the possibility of keeping the flow velocity in the second section 12 lower than would otherwise be possible with regard to the oil production, so that the pressure loss and thus the temperature drop also remain low. For some applications, this compensates for the disadvantage that the lowest temperature does not occur at the highest point of the evaporator.

It is expedient if the two evaporator sections are parts of a uniform, one-piece plate evaporator. However, a multi-part design should not be ruled out, the evaporator plate (s) forming the first section being arranged in a group in a different way and at a different location than that forming the second sections. It is important that the evaporator sections form a uniform flow path connected to a single injection valve.

It is not necessary that the two sections can be distinguished from one another by a sudden change in cross-section. Rather, a gradual transition is also conceivable.

The diagram in FIG. 4 illustrates the temperature profile of the refrigerant and in relation to the temperature profile 19 of the sprinkler water in ° C. over the height H of a plate heat exchanger according to FIG. 1 in solid lines. The refrigerant-side temperature profiles of a flooded evaporator are shown in broken lines and a conventional injection evaporator is shown in dash-dotted lines. The refrigerant and the sprinkling water move in counterflow.

The flooded evaporator achieves the most uniform temperature profile, in which, in a typical application example, the low pressure drop causes a temperature difference of only around 0.5 ° C across the height of the evaporator. On the other hand, the conventional desuperheater shows a sharp drop in temperature of, for example, 9 ° C with a risk of icing up in the middle.

The temperature profile of the evaporator according to the invention comprises a lower section 11 ', which corresponds to the lower evaporator section 11 and in which the temperature reduction corresponds approximately to that of the flooded evaporator. At the top, the second curve section 12 'follows, which corresponds to that part of the second evaporator section 12 in which the liquid phase is still present and in which the temperature accordingly drops in accordance with the reduction in the evaporation temperature caused by the pressure drop. However, since the flow path length in the narrow flow cross section is much less than in conventional injection evaporators, only a correspondingly lower pressure loss will take place overall. In addition, the lowest temperature point is close to the uppermost point of the evaporator, where the temperature of the sprinkling water is relatively high and the risk of icing is therefore low. This is followed by a curve section 12 'which corresponds to that part of the second evaporator section 12 in which the overheating of the dry, gaseous refrigerant takes place.

In summary, it can be said that in the critical, lower evaporator area, the temperature profile of the injection evaporator according to the invention is very similar to that of a flooded evaporator and that it is therefore also suitable for applications in which the temperature profile of the medium to be cooled has to be controlled, for example closely its freezing point, like this is required for the water side with temperature curve 19 in diagram 4 with cooling down to 0.5 ° C.

If, instead of the evaporator arrangement according to FIG. 1, one chooses that according to FIG. 3, the temperature in the first section remains 11 ′. For the second section, the dotted temperature curve 12 ergibt results, the lowering of the temperature of which in relation to the water curve 19 is somewhat less favorable because the temperature minimum is reached at a lower water temperature; however, this minimum is at a higher temperature than in the case of curve 12 'because the falling arrangement of the second evaporator section enables lower gas velocities and thus less pressure loss.

Claims (12)

1. A method of operating a refrigerator having a heat exchanger as an injection evaporator, in which the refrigerant from a first section of the heat exchanger of broad flow cross-section and gravity distribution of the upwardly conveyed refrigerant passes in the not yet completely evaporated state into a second section of narrower flow cross-section, wherein it has a sufficiently fluid component to prevent precipitation of lubricant, and in that in the second section, in which it is completely evaporated, the refrigerant has sufficient velocity to convey precipitated oil.

2. A heat exchanger for carrying out the method according to Claim 1, having a first section (11) of broad flow cross-section and gravity distribution of the upwardly conveyed refrigerant, and having a second section (12) of narrower flow cross-section, wherein the heat exchanger is in the form of a vertical plate evaporator.

3. A heat exchanger according to Claim 2, characterised in that the first section (11) is in the form of a flooded evaporator.

4. A heat exchanger according to Claim 2 or 3, characterised in that the first section (11) is formed with a substantially vertical flow direction.

5. A heat exchanger according to Claim 4, characterised in that the interior of the first section is divided by welded joints offset with respect to the vertical direction.

6. A heat exchanger according to any one of Claims 2 to 5, characterised in that the first section (11) occupies approximately half to three quarters of the evaporator height.

7. A heat exchanger according to any one of Claims 2 to 6, characterised in that horizontal passages (14) provided in the second portion (12) have a raised barrier (16) at their lower boundary (15) on the inlet side.

8. A heat exchanger according to any one of Claims 2 to 7, characterised in that the horizontal passages (14) forming the second portion (12) are connected to one another in ascending sequence.

9. A heat exchanger according to any one of Claims 2 to 7, characterised in that the horizontal passages forming the second evaporator portion are connected to one another in descending sequence.

10. A heat exchanger according to any one of Claims 2 to 9, characterised in that the evaporator takes the form of a trickle evaporator.

11. A heat exchanger according to any one of Claims 2 to 10, characterised in that outlet connectors (18) are connected near to the lower boundary (15) of one passage of the second section.

12. A heat exchanger according to any one of Claims 2 to 11, characterised in that it is connected to an injection valve (4) which is a thermostatic control valve with a connection (5) to a duct (7) for removing overheated refrigerant gas from the evaporator (1).

Images