US20110209485A1

US20110209485A1 - Suction superheat conrol based on refrigerant condition at discharge

Info

Publication number: US20110209485A1
Application number: US12/671,970
Authority: US
Inventors: Alexander Lifson; Michael F. Taras
Original assignee: Carrier Corp
Current assignee: Carrier Corp
Priority date: 2007-10-10
Filing date: 2007-10-10
Publication date: 2011-09-01
Also published as: CN101842646B; HK1148810A1; CN101842646A; EP2198212A4; WO2009048466A1; EP2198212A1

Abstract

A relationship is developed between a discharge condition of a refrigerant leaving a compressor, and the suction superheat or refrigerant quality. By relying upon measurement and control of the discharge condition, the present invention is able to achieve very low suction superheat values. Controlling the operation to very low suction superheat values results in augmented refrigerant system performance.

Description

BACKGROUND OF THE INVENTION

This application relates to a refrigerant superheat control to enhance system performance and improve compressor reliability, which relies upon a refrigerant thermodynamic condition at discharge to provide reliable suction superheat control.
In air conditioning, heat pump and refrigeration systems, a superheat of the refrigerant leaving an evaporator needs to be closely controlled. Refrigerant leaves the evaporator normally at the superheated thermodynamic state, where its actual temperature is higher than the corresponding saturation temperature (a superheat is defined as the difference between these two temperatures). A certain (positive) superheat is typically required to ensure that little or no liquid refrigerant enters the compressor and system operation is stable. If a significant amount of liquid refrigerant enters the compressor, an undesirable condition known as “flooding” will occur. Flooding could cause “liquid hammer” conditions damaging or breaking compression elements, dilute lubrication oil and wash it off the bearing surfaces, pump lubrication oil out of the compressor sump, and eventually degrade refrigerant system performance and operation.
On the other hand, it is known that in order to assure the highest performance (efficiency and capacity) of the refrigerant system, close to zero superheat values for the refrigerant leaving the evaporator are to be maintained. Further, by reducing suction superheat, the oil return to the compressor is also improved, as the oil is typically accumulated in the evaporator superheated section. Also, the oil viscosity is reduced with the superheat reduction due to the fact that more refrigerant is diluted in the oil at lower superheat values, and to a smaller degree, due to a saturation suction temperature increase. Conversely, as the superheat value is increased, refrigerant is boiled off from the oil increasing the oil viscosity and making the oil more susceptible to stagnate in the evaporator exit section or in the piping connecting the evaporator to the compressor. Of course, improving oil return is a goal of a refrigerant system designer, as it enhances compressor reliability and improves system performance by preventing oil retention in the evaporator and associated piping. Also, at some operating conditions, higher suction superheat values lead to elevated discharge temperatures, operational envelope reduction, potential oil breakdown and thermal distortion of compression elements.
While it is known to be desirable to reduce the superheat to the lowest value possible, to date, most refrigerant systems, at best, would operate with superheat values at the evaporator exit in a range of 5-10° F. The potential for a measurement error, due to temperature sensor measurement tolerances, calibration and resolution; system component manufacturing variability; ambient effects on system operation; load demand fluctuations and associated transient phenomena, concurrently occurring within the refrigerant system, have typically provided a practical bar to further reduction in the superheat setting.
It is undesirable, as mentioned above, to have significant flooding in the compressor, due to associated reliability issues. Thus, the refrigerant system designers have erred on the side of applying sufficient superheat values to eliminate any potential for such flooding at an entire spectrum of operating conditions. As mentioned above, uncontrolled flooding results in a drastic drop in compressor capacity and efficiency, and may also cause severe damage to the compressor.

SUMMARY OF THE INVENTION

The present invention utilizes a realization that a given change in the suction superheat will result in an expected change in a discharge temperature (or superheat) of the refrigerant leaving the compressor. That is, there is an approximately linear relationship between the suction superheat and the discharge temperature (or superheat) for the refrigerant leaving the compressor. This relationship is essentially linear at any given system operating suction and discharge pressure.
Thus, by monitoring the discharge temperature (or superheat) and changing/controlling a condition of the refrigerant leaving the evaporator or entering the compressor (superheat or quality) based on discharge temperature, the system can be reliably operated at a desired low suction superheat or have a minimal controlled amount of liquid refrigerant entering the compressor suction port. The control of the suction superheat that is based on discharge temperature (or superheat) can, for example, be accomplished by varying the opening of an expansion valve or a suction modulation valve.
The relationship between the discharge temperature (or superheat) and suction superheat can be determined experimentally or can be developed analytically. Further, the relationship can be periodically tested/verified during operation to ensure that the relationship still holds, or any small adjustments need to be made to this relation based on these periodic tests.
In disclosed embodiments, the present invention allows for operation at the superheat levels that are substantially lower than the superheat levels in the 5 to 10° F. range reliably achievable in the past. With the current technique, the superheat levels leaving the evaporator can be as low as 1 to 2° F., or with appropriate control, some minimal amount of liquid refrigerant can be allowed to enter the compressor suction port.
These and other features of the present invention can be best understood from the following specification and drawings, the following of which is a brief description.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a view of a refrigerant system incorporating the present invention.

FIG. 2 is a chart showing changes in discharge temperature as a function of suction superheat.

FIG. 3 shows the system efficiency with respect to suction superheat and discharge temperature.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

A basic refrigerant system 20 is illustrated in FIG. 1 and incorporates a compressor 22 delivering compressed refrigerant downstream to a heat rejection heat exchanger 24 (a condenser for subcritical applications and a gas cooler for transcritical applications). An expansion device 26 is preferably an electronic expansion device, and is generally known in the industry. Refrigerant having passed through the expansion device 26 flows in sequence through an evaporator 28, through an optional suction modulation valve 30, and through a suction line 38 back to the compressor 22. As is shown in this Figure, a temperature sensor 46 is placed on or inside the discharge line leaving the compressor. The temperature sensor 46 can also be positioned to measure the discharge temperature on the compressor shell or inside the compressor shell. The temperature sensor 46 communicates with an electronic controller 32, which in turn controls the electronic expansion device 26, or/and the optional suction modulation valve 30 to adjust and control the suction superheat. There is also a sensor 56 measuring the suction superheat value that is placed between the expansion device 26 and compressor 22. More preferably, the sensor 56 is placed on the suction line 38 between the exit from the evaporator 28 and inlet to the compressor 22. The sensor 56 can be a sensor that directly measures superheat (the difference between the actual and saturated refrigerant temperatures at approximately the same location). Alternatively, the sensor 56 can be a temperature sensor. The temperature sensor 56, for example, can be of a thermocouple or thermistor type. In the latter case of the temperature sensor 56, an additional temperature sensor 58 may be placed within the evaporator 28 (in the refrigerant flow or externally on the evaporator surface) within the two-phase region to determine the saturated refrigerant temperature in the evaporator. The difference between these two temperature measurements provided by the temperature sensors 56 and 58 would determine the suction superheat value. Instead of using the temperature type sensor 58 within the evaporator, a pressure type sensor 60 may be utilized that would determine pressure of the refrigerant at or near the location where the superheat value is to be obtained (e.g. suction side of the refrigerant system 20). After the refrigerant pressure value is measured by the pressure sensor 60, it can be converted to a corresponding saturated temperature value, as known. The suction superheat value is then simply calculated by subtracting the obtained saturated refrigerant temperature from the measured actual refrigerant temperature. As mentioned above, the temperature sensors 56 and 58 can be installed on the external surface (e.g. air side) of the tubing, compressor, heat exchanger, etc. These sensors can also be installed internally or within the refrigerant flow (so-called in-flow sensors). In the latter case, the in-flow sensors would measure the temperature of the refrigerant directly. If the temperature sensors are installed externally, they are preferably insulated or shielded from the ambient environment to reduce the measurement error.
The present invention allows achieving very low superheat values or controlled flooded conditions for the refrigerant prior to entering the compression chambers by relying upon a relationship that is illustrated in FIGS. 2. As shown in FIG. 2, the discharge temperature as a function of suction superheat is plotted for a particular operating condition. Further, data presented in FIG. 3 shows the system thermodynamic efficiency in relation to suction superheat and discharge temperature. As can be seen from FIG. 2, a change in suction superheat essentially results in a linear change in the discharge temperature until the compressor begins to experience a flooded condition. The flooded condition is represented by a vertical line originating at the point “O” and extending downward. As the amount of flooding increases, the discharge temperature continues to decrease, while the suction superheat value remains at zero (when the superheat is zero, the actual temperature of the refrigerant is equal to the saturated temperature). As shown in FIG. 2 and FIG. 3, the most efficient compressor, evaporator and the entire refrigerant system operation is achieved in the region located between the points “C” and “G” shown on the graphs. The point “C” corresponds to approximately 4° F. of suction superheat and point “G” corresponds to slight compressor flooding conditions.
In this example, if the compressor operation is controlled based on the discharge temperature, and at 160° F. in particular, it will correspond to a condition that falls approximately in the middle of the most efficient region of the operation, the point “E”. The point “E” is located approximately in the middle of the region defined by the points “C” and “G”. Further, the measurements error tolerance band for the discharge temperature is defined in the FIGS. 2 and 3 to be between the points “D” and “F”. Since the measurement error tolerance band falls within the region of the most efficient operation, it can be concluded that, by controlling the suction superheat based on the discharge temperature, the refrigerant system 20 can always operate at or near the most efficient point. Further, in this particular example, the refrigerant system 20 can be comfortably operated at the point “E” using the discharge temperature control, where the point “E” corresponds to only 1° F. of suction superheat.
On the other hand, as has been done in the prior art, if the control of suction superheat is executed by direct measurements of the suction superheat value, the suction superheat has to be set to at least 5° F., as shown in FIG. 2 by the point “B”, to make sure that an uncontrolled flooding situation has not occurred (uncontrolled flooding would represent large amounts of liquid refrigerant reaching the compressor suction port). For example, if the superheat value is set below 5° F., then with potential measurement errors, the actual operating point could be well bellow the point “H”, which corresponds to severe flooding conditions for the compressor, and its potential damage. Thus, for this example, the minimum acceptable setting for the suction superheat using the prior art technique would be a value of 5° F.
As can be seen from FIG. 3, operation at the point “B” is not nearly as efficient as operation at the point “E”. For example, for typical frozen mode operating conditions in the truck-trailer or marine container applications, it would translate into 2% drop in the system performance (capacity and efficiency) when the suction superheat value is increased from 1 to 5° F. It should be pointed out that, due to measurement error and reliability concerns, the refrigerant system performance reduction can be substantially higher, if for example, the actual superheat value to be set at the point “A”, if the prior art technique is employed. Stated differently, the present invention allows the use of a discharge temperature sensor to precisely control suction superheat by utilizing a relation that defines suction superheat value based on the measurements of the discharge temperature. For example, in using this relation, if the discharge temperature is set at 162° F. the refrigerant system can be operated at 1° F. suction superheat without any danger of reaching the point “H” that corresponds to the beginning of the region for severe flooding conditions that can damage the compressor. The controller 32 for the refrigerant system 20 can utilize the sensed discharge temperature to predict an amount by which the discharge temperature must be changed to achieve a desired suction superheat value. In disclosed embodiments, the controller 32 would then control the expansion device 26, and/or the suction modulation valve 30, to change the discharge temperature, and hence adjust the suction superheat of the refrigerant reaching the compressor. The suction superheat can be lowered to the values below 2° F., in the disclosed embodiments, and may be kept at approximately 1° F.
The refrigerant system can be periodically tested, to assure that the FIG. 2 relationship holds, by raising the suction superheat from a lower level to a higher level, and then measuring an inter-relation between the change in the discharge temperature and the suction superheat, to ensure that the expected change is still taking place. This can occur at high enough values of suction superheat so that the suction temperature can be reliably measured. As an example, the suction superheat can be changed from 1 to 16° F. and the corresponding discharge temperature changes can be measured at the 16° F. superheat region. Therefore, the refrigerant system may operate in the self-learning or adaptive manner to achieve the highest operational performance possible.
While the graph presented in FIG. 2 shows the relation between discharge temperature and suction superheat for a particular suction and discharge pressures, similar graphs can be developed for other operating conditions. These graphs then can be used to control the suction superheat at any expected operating conditions. Instead of developing multiple graphs, the results can be assembled into look-up tables, and then interpolated as needed for actual values of suction and discharge pressures. Furthermore, instead of using the look-up tables or graphs, equations relating suction superheat to discharge temperature for a given suction and discharge pressures can be developed.
As mentioned above, the refrigerant system can be self-learning such that, during the system operation, for a given suction and discharge pressures, the discharge temperature can be varied on an intermittent basis to establish a relation between the discharge temperature and suction superheat. Stated differently, the refrigerant system can itself develop the graph of FIG. 2 during operation. Then the corresponding graph values can be stored in the memory of the refrigerant system controller as a function of suction and discharge pressures, and then retrieved by the controller from the memory on an as needed basis.
Furthermore, since the discharge side pressure or saturation temperature may be known, a similar relationship can be established between the suction and discharge superheat that can be used for identical purposes of the refrigerant system control.
As mentioned above, previous attempts to use suction and discharge pressures and discharge temperature to reliably control suction superheat to extremely low values have failed, since they relied on refrigerant properties and compression process polytropic exponent both of which highly depend on operating conditions and compressor design characteristics. This becomes particularly difficult for the compressors with a built-in volume ratio that are subjected to over-compression or under-compression conditions. Therefore, the prior art methods could be used as the first order approximations only and could not be relied upon to control suction superheat to near zero values.
It should be pointed out that many different compressor types could be used in this invention. For example, scroll, screw, rotary, or reciprocating compressors can be employed.
The refrigerant systems that utilize this invention can be used in many different applications, including, but not limited to, air conditioning systems, heat pump systems, marine container units, refrigeration truck-trailer units, and supermarket refrigeration systems.
Although a preferred embodiment of this invention has been disclosed, a worker of ordinary skill in this art would recognize that certain modifications would come within the scope of this invention. For that reason, the following claims should be studied to determine the true scope and content of this invention.

Claims

1. A refrigerant system comprising:

a compressor, said compressor having a suction inlet line and a discharge outlet line;

a compressed refrigerant passing from said compressor downstream to a heat rejection heat exchanger and then downstream to an expansion device;

an evaporator positioned downstream of said expansion device; and

a sensor for sensing a condition of a discharge refrigerant leaving the compressor and

a control utilized to control a refrigerant thermodynamic state at a location between the expansion device and the compressor based upon the sensed condition of the discharge refrigerant.

2. The refrigerant system as set forth in claim 1, wherein said sensed condition is the discharge temperature of the refrigerant.

3. The refrigerant system as set forth in claim 2, wherein said refrigerant thermodynamic state is at least one of the suction superheat or refrigerant quality.

4. The refrigerant system as set forth in claim 1, wherein said sensed condition is the discharge superheat of the refrigerant.

5. The refrigerant system as set forth in claim 1, wherein said location is between the evaporator and the compressor.

6. The refrigerant system as set forth in claim 5, wherein a relationship is found between at least one of the discharge temperature and discharge superheat, and at least one of the suction superheat and refrigerant quality at a location between the evaporator exit and the compressor inlet.

7. The refrigerant system as set forth in claim 6, wherein said relationship is essentially linear.

8. The refrigerant system as set forth in claim 6, wherein said relationship is periodically checked by varying said suction superheat value during refrigerant system operation.

9. The refrigerant system as set forth in claim 6, wherein said relationship for at least one of suction pressure and discharge pressure is stored in the system controller memory.

10. The refrigerant system as set forth in claim 9, wherein a control retrieves said relationship from said memory.

11. The refrigerant system as set forth in claim 9 wherein said relationship is a lookup table.

12. The refrigerant system as set forth in claim 9, wherein said relationship is determined experimentally.

13. The refrigerant system as set forth in claim 12 wherein said relationship is determined during unit operation.

14. The refrigerant system as set forth in claim 9, wherein the relationship is determined analytically.

15. The refrigerant system as set forth in claim 1, wherein the sensed condition is used to achieve suction superheat values equal to or less than 2° F.

16. The refrigerant system as set forth in claim 1, wherein the sensed condition is used to achieve suction superheat values equal to or less than 4° F.

17. The refrigerant system as set forth in claim 5, wherein said suction superheat is calculated based on the difference between the saturated temperature and measured temperature sensed between the evaporator and compressor.

18. The refrigerant system as set forth in claim 17, wherein said suction temperature is measured by a temperature sensor.

19. The refrigerant system as set forth in claim 18 wherein said sensor is one of the thermistor or thermocouple type.

20. The refrigerant system as set forth in claim 17, wherein said saturated temperature is measured by a temperature sensor located within the evaporator.

21. The refrigerant system as set forth in claim 17, wherein said saturated temperature is calculated based on the pressure measurement.

22. A method of operating a refrigerant system including the steps of:

providing a compressor, said compressor having a suction inlet line and a discharge outlet line;

passing a compressed refrigerant from said compressor downstream to a heat rejection heat exchanger and then downstream to an expansion device;

passing the refrigerant to an evaporator downstream of said expansion device;

sensing a condition of a discharge refrigerant leaving the compressor; and

controlling a refrigerant thermodynamic state at a location between the expansion device and the compressor based upon the sensed condition of the discharge refrigerant.

23. The method as set forth in claim 22, wherein a relationship between said refrigerant thermodynamic state at a location between the expansion device and the compressor and said sensed condition of the discharge refrigerant used for the control is determined experimentally.

24. The method as set forth in claim 22, wherein a relationship between said refrigerant thermodynamic state at a location between the expansion device and the compressor and said sensed condition of the discharge refrigerant used for the control is determined analytically.

25. The method as set forth in claim 22, wherein the sensed condition is used to achieve suction superheat values equal to or less than 2° F.