EP0356642B1 - Thermostatic expansion valve - Google Patents

Description

The invention relates to a thermostatic expansion valve for a refrigeration system with a compressor, a condenser and an evaporator, wherein the refrigerant liquefied in the condenser flows through a throttle point of the expansion valve and the opening width of the throttle point is determined by the position of a throttle body, and the position of the throttle body can be influenced by a pressure and / or temperature sensor or a membrane, the Throttle body facing side is acted upon by a control medium, which acts in an opening sense on the throttle body, while a spring acts in a closing sense on the throttle body, furthermore the overheating value (Δtü) of the refrigerant after the evaporator can be detected and via one sensor the signal corresponding to the overheating setpoint (Δtü / Soll) can be acted on the throttle body in a regulating sense via a controllable device.

Such a thermostatic expansion valve is known from DE-A-27 49 250. It is used in conjunction with a conventional refrigeration system, with the signals for regulating the superheating temperature being taken directly from the coolant circuit to a predetermined target value. For this purpose, there are sensors in front of and behind the evaporator, which convert the determined temperature value into an electrical signal, which is transmitted to a control unit. The latter then in turn has a corresponding regulating action on the throttle body in the sense of transferring the instantaneous actual value to the target value.

A similar thermostatic expansion valve has become known from US Pat. No. 4,750,334, which is likewise inserted into a coolant circuit which is the same as that of the present expansion valve or that of DE-A-27 49 250 corresponds. The temperature is only recorded at the outlet of the evaporator. This measured value is also used for control purposes, the throttle body being adjusted so that the actual value of the superheating temperature is brought up to the target value.

A further thermostatic expansion valve of this type is shown in FIG. 1 of the drawing of the subject matter of the invention. Here, refrigerant vapor (suction gas) is drawn in by a compressor 1 at low pressure and low temperature and brought to a higher pressure. The refrigerant is heated. This is followed by condensation in condenser 2 at high pressure with heat being released to the environment. The liquefied refrigerant is then passed through the throttle 3 of the expansion valve 4. The throttling leads to a decrease in pressure and temperature with partial evaporation of the liquid refrigerants. The refrigerant then passes from the throttling point 3 to the evaporator 5. There, the refrigerant is removed from the environment to be cooled, e.g. the interior of the motor vehicle, heat is supplied. The remaining part of the refrigerant is evaporated. The cycle then begins anew.

When the suction gas is returned from the evaporator 5 to the compressor 1, the refrigerant is passed through the suction gas space 6 of the expansion valve 4. Is located above the suction gas chamber 6 the control head 7, in which a membrane 8 is arranged. Its lower surface has thermal contact with the suction gas space 6 and thus with the refrigerant vapor flowing from the evaporator 5 to the compressor 1. The control head 7 is filled with control medium 9 above the membrane 8. This is usually identical to the refrigerant in the circuit. The control medium is designed so that it is present as an overheated gas under operating conditions with evaporation temperatures below approx. + 10 ° C as a wet steam mixture, above approx. 15 ° C at maximum operating pressure. A transmission rod 10 is connected to the membrane 8. The throttle body 11 is fastened at its lower end. Its relative position in relation to the throttle point 3 determines the cross-section through which it flows and thus the degree of throttling of the refrigerant. The liquefied refrigerant coming from the condenser 2 first enters the room 26 (high pressure side) and from there via the throttle 3 in the room 29 (low pressure side). A spring 12 acts on the throttle body 11 from below, the preload of which can be adjusted by means of a spindle 13. The transmission rod 10 is sealed by means of a seal 14, the spindle 13 by means of a seal 15 in the housing. If a certain operating state is reached, the force resulting from the pressure of the control medium 9 in the control head 7 is on the membrane 8, the force exerted on the membrane 8 by the pressure of the suction gas in the suction gas 6, and by the spring 12 on the Throttle body 11 exerted force in equilibrium.

The aim of setting the working conditions of the expansion valve 4 is to work with a relatively large "overheating" of the suction steam behind the evaporator 5 and thus also in the suction gas space 6 for reasons of the effectiveness of the cooling capacity. This is necessary, among other things, because part of the liquid refrigerant remains dissolved in the oil added to lubricate the compressor and thus does not contribute to cooling. This proportion of refrigerant dissolved in the oil decreases with increasing overheating. On the other hand, the refrigerant behind the compressor 1 should not exceed a maximum temperature tmax of, for example, 150 ° C. On average, the temperature should not be higher than 130 ° C, for example. Overheating is the temperature difference between the refrigerant vapor and the refrigerant present as a wet vapor mixture at the same pressure. For clarification, reference is made to the enthalpy (log p / h) diagram according to FIG. 2. Curve K denotes the boiling or dew line, ie it includes the wet steam range. The lines AB or A'B ', BC or B'C, CD and DA or CA' denote the changes in state in the compressor 1, in the condenser 2, in the expansion valve 4 and in the evaporator 5. For clarification, some temperature values have been entered. In the case of a design characterized by points A and B, the superheating Δtü after the evaporator 5 is, for example, 5 ° C. Under the assumed operating conditions, this leads to the refrigerant vapor behind the compressor 1 having the maximum permissible Temperature tmax of 150 ° C reached. In another design, which is characterized by points A 'and B', the overheating after the evaporator 5 is, for example, 3 ° C., with the result that behind the compressor 1 only the temperature of the refrigerant vapor which is tolerable as the average value of approximately 130 ° C is reached. To complete the figure, the pressure loss P _EX which occurs when it passes through the expansion valve 4 and the pressure loss P _V which occurs in the evaporator 5 are also entered.

The problem with refrigeration systems in motor vehicles is that the operating conditions change constantly; the output of the compressor, which is driven by the motor of the motor vehicle, changes continuously with the speed: In FIG. 3, typical operating states, namely operating states 64/2, 32/2 and IT, are shown in the enthalpy diagram. These are the acceptance values common in automotive engineering, in which maximum cooling capacity = minimum interior temperature is required. IT (idle test) is an operating state of a motor vehicle air conditioning system in which the vehicle is at a standstill and the compressor is driven by the vehicle engine running at idling speed. The ventilation of the condenser is the most unfavorable in this case, so that particularly high pressures and hot gas temperatures can occur. In operating mode 32/2, the driving speed is 32 km / h, the gear shift position is second gear. This is an im general uncritical driving condition. 64/2 (64 km / h; 2nd gear) is a driving condition in which the most critical hot gas temperatures currently occur.

The dash-dotted curve results from FIG. 2 as a typical course of the overheating Δtü. In the operating states IT and 64/2, the maximum permissible temperature of the refrigerant vapor tmax is reached. This is not the case in operating state 32/2. Here at B 'only a lower temperature is reached behind the compressor. In order to increase the effectiveness of the refrigeration system in this operating state, it would be desirable and also possible to work with a higher superheat, to the extent that B 'would also lie on the curve tmax. The dash-dotted line results from these requirements as the desired characteristic curve Δtü / Soll. To put it simply, the requirement can be expressed that the overheating should be around 8 K for reasons of effectiveness, but that it should be controllable down to 1 to 2 K in certain operating conditions in order to avoid inadmissibly high refrigerant vapor temperatures and to avoid early evaporator icing.

The object is to develop a thermostatic expansion valve of the type described in the introduction in such a way that the superheating Δtü of the refrigerant vapor is regulated according to the Evaporator depending on the operating conditions of a motor vehicle in which the expansion valve is installed or depending on new criteria of the known coolant circuit is possible such that the maximum permissible temperature tmax after the compressor is not exceeded.

This object is achieved with the features of claim 1. Advantageous further developments are defined in the subclaims.

Figur 1

den Kreislauf einer Kälteanlage nach dem vorbekannten Stand der Technik;

Figur 2

das Enthalpie-Diagramm für den Kreislauf nach Figur 1;

Figur 3

verschiedene Betriebszustände eines Kraftfahrzeugs im Enthalpie-Diagramm;

Figuren 4 bis 8

5 Ausführungsbeispiele

   Soweit die Teile in den Figuren 4 bis 8 nicht besonders gezeichnet sind, sind sie dieselben wie in Figur 1; außerdem sind in den Figuren 4 bis 8 Verdichter Kondensator und Verdampfer der Einfachheit halber weggelassen.Embodiments of the invention and its advantageous developments are described below. The drawings show:

Figure 1

the circuit of a refrigeration system according to the prior art;

Figure 2

the enthalpy diagram for the circuit of Figure 1;

Figure 3

different operating states of a motor vehicle in the enthalpy diagram;

Figures 4 to 8

5 working examples

As far as the parts in Figures 4 to 8 are not specially drawn, they are the same as in Figure 1; in addition, in FIGS. 4 to 8, the condenser and evaporator are omitted for the sake of simplicity.

In the exemplary embodiment according to FIG. 4, the spring 12, which presses the throttle body 11 in the closed position, is supported on an adjusting plate 20 which is arranged in a pressure cell 21 so as to be displaceable in height. The space 22 below the adjusting plate 20 is connected in terms of pressure to the suction gas space 6 via a line 23. The space 24 above the adjusting plate 20 is connected via a line 25 to the space 26 (high pressure side) into which the liquefied refrigerant flows from the condenser. A further spring 27 acts on the adjusting plate 20 Pressure on the adjusting plate 20 is adjustable by means of the setting wheel 28. In addition to the force of the springs 12, 27, the adjusting plate 20 is therefore also subject to the differential pressure between the pressure in the space 26 (high pressure side) and the pressure in the suction gas space 6 (low pressure side). Since the decisive pressure drop takes place in the entire refrigerant circuit at the throttle point 3 of the expansion valve 4, the pressure in the space 26 is greater than the suction gas space 6. The differential pressure is therefore in the opposite direction to the pressures of the two springs 12, 27 on the adjusting plate 20 on the throttle body 11 in Opening direction of throttle 3 effective. As a result:

If a too high pressure of the liquefied refrigerant entering the space 26 of the expansion valve 4 occurs behind the compressor 1 and thus also behind the condenser 2, which is associated with an excessively high temperature behind the compressor 1, this leads to a corresponding pressure increase in the room 24 and thus to relieve the springs 12, 27. The throttle body 11 thus moves downward. The cross section of the throttle point 3 is enlarged. An increased mass flow flows through the throttle point 3. This leads to a reduction in the overheating Δtü in the suction gas space 6. The same effect occurs when the pressure in the suction gas space 6 drops. The system is designed in such a way that at evaporation temperatures of 0 ° C with optimal performance overheating Δ tü. This results in overheating in the IT operating state with driving _values of t (p _RVa ) greater than or equal to 0 ° C (R: refrigerant; V: evaporator; a: exit). It is therefore possible to use this embodiment with maximum operating pressure ( M aximum O perating P ressure = MOP). In addition, the changes required for expansion valves currently in series production are minor.

In the exemplary embodiment according to FIG. 5, the spring force is changed by an electrically controllable control element. The spring 12, which presses the throttle body 11 in the closing direction, is mounted on a platform 30, which is arranged on the top of a control piston 31, which can be moved up and down in a pressure cell 32. The control piston 31 is pressed upwards by a spring 33. The pressure cell 33 is connected via the connecting piece 34 to a cylindrical space 35 in which a servo piston 36 swings back and forth. The servo piston 36 is provided with two collars 37, 38. The high pressure in front of the throttle point 3 acts on the left collar 37 via opening 43, space 44 and opening 45. A further spring 39 acts on the right collar 38 of the servo piston 36 and is supported on the housing. The left end of the servo piston 36 is connected to the tappet 40 of a solenoid valve 41. The space to the right of the collar 38, in which the spring is arranged, is connected via the opening 46, the space 47 and the line 48 to the low pressure in space 29 behind the throttle 3. When the solenoid valve 41 is de-energized, the high pressure of the liquefied refrigerant coming from the condenser 2 acts on the servo piston to the left of the collar 37 and to the right of the collar 38 the spring 39. The servo piston 35 swings back and forth, whereby in one end position the collar 37 opens the opening 45 and in the other end position the collar 38 briefly opens, the other opening to the room 35 then being closed. Accordingly, a further, smaller refrigerant partial flow takes place intermittently parallel to the (main) throttle point 3 via the openings 45 and 46 and space 35. This partial refrigerant flow is interrupted when the solenoid valve 41 is energized and fixed in its left end position, in which the left collar 37 connects the opening 45 to the space 35 between the two collars 37, 38. Then the pressure of the high-pressure side of the expansion valve 4 can propagate in the space 26 into the interior of the pressure cell 32 and thus presses the control piston 31 in addition to the springs 12, 33 upwards. The cross section of the throttle 3 is thus reduced. Accordingly, the mass flow through the throttle point and thus the cooling is reduced, so that the overheating door is increased. By releasing the pressure in the pressure can 32 (control pressure), intermediate positions are also possible. The solenoid valve 41 can be clocked as a function of a measurement of the temperature downstream of the compressor. With appropriate clocking, this exemplary embodiment also enables a sliding overheating setting between two limit values. It is characterized by a high response speed. The magnetic valve 41 only has a servo function, and can therefore be made correspondingly small. With a flushing facility and a comparatively large size There is practically no risk of clogging in cross-sections in the connecting channels. In the event of electrical faults, the valve works with slight overheating. This results in good emergency running properties.

The exemplary embodiment according to FIG. 6 works with a thermomotor 60 as a control element, which is formed by a control medium 62 in a corrugated bellows 61 and a heating plate 63, which can be heated via a connection 64, and is surrounded by an insulation 65. A further spring 66 for high pressure compensation is provided in the corrugated bellows 61. The thermal motor 60 is located inside the insulation 65 in a can 67, which is vented via line 68.

The thermal motor 60 acts on an adjusting plate 70 which is arranged such that it can be displaced in height between the seals 71 and which in turn supports the spring 72. The spring 72 presses with its upper end against the throttle body 11, in addition to the spring 12, which is supported on a platform 30 which is seated on the collar 50. Depending on the heating power that is supplied to the heating plate 63, the control medium 62 expands in the corrugated bellows 61 and thus presses the adjusting plate 70 upwards and thus increases the pressure of the spring 72 on the throttle body 11 . In the exemplary embodiment shown in FIG. 6, it is also advantageous that the thermal motor 61 is connected via line 68 to the cooled refrigerant in the space 29 behind the throttle 3 stands, so that the thermal motor can be cooled in this way. The heating plate 63 can preferably be implemented by an intrinsically safe PTC resistor. The corrugated bellows construction is suitable with regard to the required travel, which in practice is 1.5 to 2 mm. The exemplary embodiment according to FIG. 6 is a particularly simple mechanical system with small time constants. If the heating fails, reliable emergency running properties result, since the expansion valve 4 then works with low overheating (throttle 3 far open). A further variation (not shown) of the exemplary embodiment according to FIG. 6 could consist in that the thermal motor is also placed in the suction gas space 6 because of the cooling and the power transmission to the throttle body 11 is realized by a lever linkage.

bei ca. 3 bar Verdampfungdruck zur Abregeglung der Überhitzung auf 2 K eine Leistung von ca. 3 W erforderlich ist. Sie setzt sich zusammen aus der für die Zustandsänderung der Steuerfüllung 9 notwendigen Leistung und dem Wärmeverlust über der Membranfläche 8. Einschließlich weiterer Verluste an die Umgebung kann mit einer Leistung unter 20 W gerechnet werden. Wichtig ist, daß das Ventilgehäuse in seiner Gesamtheit aus elektrisch schlecht leitendem Werkstoff ausgebildet sein sollte, um die Effektivität der Einwirkung der Induktionsspule 80 auf die Steuermembran 8 zu erhöhen.In the exemplary embodiment according to FIG. 7, the control head 7 is surrounded by an induction coil 80. In this way, more heat is supplied to the control head when the induction coil is excited than would be necessary per se to control the overheating Δtü. This results in an increase in pressure of the control filling 9 above the diaphragm 8, which pushes the unit formed by the throttle body 11 and the transmission rod 10 downward, so that this results in an increased mass flow and, consequently, better cooling and thus less overheating Δtü. In a practical implementation of this embodiment, it was found that, starting from one Basic setting of $\text{Δ tü = 8 K}$

at an evaporation pressure of approx. 3 bar to reduce the overheating to 2 K, a power of approx. 3 W is required. It is composed of the power required for changing the state of the control filling 9 and the heat loss over the membrane surface 8. Including further losses to the environment, a power below 20 W can be expected. It is important that the valve housing as a whole should be made of poorly electrically conductive material in order to increase the effectiveness of the action of the induction coil 80 on the control membrane 8.

The embodiment according to Figure 8 shows the direct heating of the control medium 9 by the same above the control head 7 in a cup-like recess disposed electric heating plate 90 with terminal 91, by a PTC resistor (ositive P T emperature C oefficient) or a Peltier element can be formed and is surrounded by an insulating cap 92. In this arrangement, an electrical heating power of the heating plate 90 of approximately 8 W is required in order to reduce the overheating Δtü from 7 K to approximately 1.5 K. Since in this exemplary embodiment only the control head has to be changed compared to conventional expansion valves (cf. FIG. 1), the exemplary embodiment is particularly suitable for retrofitting existing refrigeration systems. This exemplary embodiment is particularly cost-effective and, because of the low electrical heating line required, represents only a very low load on the vehicle electrical system.

• (a) die Heißgastemperatur, ermittelt durch Temperatursensoren in oder an der Hochdruckleitung vom Verdichter zum Kondensator bzw. im oder am Verdichtergehäuse selbst;
• (b) die Drehzahl des Verdichters, ermittelt z.B. aus der Motordrehzahl oder mit Hilfe eines Drehzahlaufnehmers am Verdichter
• (c) die Oberflächentemperatur des Verdampfers ermittelt z.B. durch Temperatursensoren im Verdampfernetz bzw. in oder an der Leitung vom Verdampfer zum Verdichter;
• (d) die Lufttemperatur nach Verdampfer ermittelt durch einen Temperatursensor im Luftstrom nach Verdampfer.

The criteria for controlling the valves (FIG. 5) or heating elements (FIGS. 6 to 8) are:

• (a) the hot gas temperature, determined by temperature sensors in or on the high-pressure line from the compressor to the condenser or in or on the compressor housing itself;
• (b) the speed of the compressor, determined, for example, from the engine speed or with the aid of a speed sensor on the compressor
• (c) the surface temperature of the evaporator is determined, for example, by temperature sensors in the evaporator network or in or on the line from the evaporator to the compressor;
• (d) the air temperature after the evaporator is determined by a temperature sensor in the air flow after the evaporator.

Claims (6)

1. A thermostatic expansion valve (4) for a refrigerating installation with a compressor (1), a condenser (2) and an evaporator (5), the refrigerant liquefied in the condenser (2) flowing through a throttling means (3) in the expansion valve (4), the size of the opening in the throttle means (3) being determined by the position of a throttle member (11), the position of the throttle member (11) being capable of being influenced by a pressure and/or temperature indicating means or a diaphragm (8), of which the side directed away from the throttle member (11) is subject to the action of a controlling medium (9) which acts on the throttle member (11) in opening direction while a spring (12) acts on the throttle member in closing direction, whereby furthermore the overheating value (Δtu) of the refrigerant downstream of the evaporator (5) being detectable by at least one sensor, a signal corresponding to the desired overheating value (Δtu desired) acting on the throttle member (11) in a stepping-down sense by means of an actuable device (36, 41; 60, 63; 80; 90), characterised in that the expansion valve (4) is installed in a motor vehicle and in that the signal for actuating the device (36, 41; 60, 63; 80; 90) is derived from the temperature of the heated gas downstream of the compressor (1), the rotary speed of the motor vehicle engine or of the compressor (1) or from the temperature of the stream of cold air downstream of the evaporator (5) and in that the side of the diaphragm (8) which is directed towards the throttle member (11) is exposed to the coolant vapour between evaporator (5) and compressor (1).
2. A thermostatic expansion valve according to Claim 1, characterised in that the spring (12) is biased on a movably disposed adjusting plate (20) one side (22) of which is subject to the refrigerant vapour between evaporator (5) and compressor (1) while its other side is exposed to the liquefied refrigerant downstream of the condenser (2).
3. A thermostatic expansion valve according to Claim 1, in which the means of stepping-down the overheating (Δtu) of the refrigerant vapour downstream of the evaporator is constituted by an electronically actuable regulating device (36, 41; 60, 63) for influencing the initial tension of the spring (12), characterised in that the spring (12) is seated on a control piston (31) which is subject to the pressure in a space (35) which can by means of a servo-piston (36) cause to communicate either through a first aperture (45) with the high pressure side (26) of the expansion valve (4) or via a second aperture (46) with the low pressure side (29) thereof.
4. A thermostatic expansion valve according to Claim 3, characterised in that the servo-piston (36) is on its side (37) and via the said first aperture (45) exposed to the pressure of the liquefied refrigerant while on its other side (38) it is subject ot the action of a spring (39) so that with alternate opening up of one of the two apertures (45, 46), it (36) performs a continuous oscillation, and in that a partial flow of liquefied refrigerant flows through the space (35), the regulating member being a magnetic valve (41) by which the servo-piston (36) can be locked.
5. A thermostatic expansion valve according to Claim 1, in which the device for stepping-down the overheating ( tu) of the refrigerant vapour downstream of the evaporator is constituted by an electronically actuable regulating means (36, 41; 60, 63) for influencing the initial tension (12) of the spring (12), characterised in that the means of stepping-down the desired value of overheating is constituted by a thermomotor (60) which supports an adjusting plate (70) for the spring (12) and is capable of being actuated via an electrically-operated heating element (63).
6. A thermostatic expansion valve according to Claim 5, characterised in that the thermomotor consists of an undulating bellows (61) filled with a controlling medium (62).

Images