EVAPORATOR REFRIGERANT SATURATION DEMAND DEFROST

Abstract

A method is disclosed for controlling initiation of a defrost cycle of an evaporator heat exchanger of a refrigeration system operatively associated with a refrigerated transport cargo box. The method includes the steps of establishing an return air-saturation temperature differential equal to the difference of a sensed air temperature of an air flow returning from the cargo box to pass over the heat exchange surface of the evaporator heat exchanger minus a refrigerant saturation temperature of a flow of refrigerant passing through the evaporator heat exchanger, comparing the return air-saturation temperature differential to a set point threshold defrost temperature differential, and if the return air-saturation temperature differential exceeds the set point threshold defrost temperature differential, initiating a defrost cycle for defrosting the evaporator heat exchanger.

Description

CROSS REFERENCE TO RELATED APPLICATION

This application claims priority to U.S. Provisional Patent Application Ser. No. 61/360,651 entitled “Evaporator Refrigerant Saturation Demand Defrost,” filed on Jul. 1, 2010. The content of this application is incorporated herein by reference in its entirety.

FIELD OF THE INVENTION

This invention relates generally to refrigeration systems and, more particularly, to refrigerant vapor compression system evaporator coil defrost control and, more specifically, to the on demand initiation of a defrost cycle for the evaporator coil in response to a differential between return air temperature and evaporator refrigerant saturation temperature.

BACKGROUND OF THE INVENTION

Refrigerant vapor compression systems are well known in the art and commonly used for conditioning air to be supplied to a climate controlled comfort zone within a residence, office building, hospital, school, restaurant or other facility. Refrigerant vapor compression systems are also commonly used in refrigerating air supplied to display cases, merchandisers, freezer cabinets, cold rooms or other perishable/frozen product storage area in commercial establishments. Refrigerant vapor compression systems are also commonly used in transport refrigeration systems for refrigerating air supplied to a temperature controlled cargo space of a truck, trailer, container or the like for transporting perishable/frozen items by truck, rail, ship or intermodal.

Refrigerant vapor compression systems typically include a compressor, a condenser, an evaporator, and an expansion device. These basic components are interconnected by refrigerant lines in a closed refrigerant circuit, arranged in accord with known refrigerant vapor compression cycles. The expansion device is disposed upstream, with respect to refrigerant flow, of the evaporator and downstream of the condenser. The evaporator includes a heat exchanger, typically a heat exchange tube coil, finned or un-finned, through which refrigerant flowing through the refrigerant circuit passes in heat exchange relationship with air drawn from and circulated back to a temperature controlled space. Because the air within the temperature controlled space will contain moisture, to varying degrees, whether the climate controlled be an air conditioned room, a refrigerated display case, or a temperature controlled transport cargo box, and because the temperature of the refrigerant flowing through the evaporator heat exchange tube coil may drop below the freezing point of water, in some applications and under certain operating conditions, moisture in the air flowing over the heat exchange tube coil will condense on the heat exchange surface of the tube coil and form frost. As the frost accumulates over time of system operation in a cooling mode, the frost builds up on heat exchange surface of the tube coil, adversely impacting heat transfer performance and restricting air flow over the tube coil.

Consequently, it is customary practice to periodically interrupt system operation in a cooling mode and enter a defrost mode wherein the accumulated frost is melted off the evaporator tube coil. A defrost cycle can be accomplished by reversing the flow of refrigerant through the refrigerant circuit so as to circulate a heated refrigerant, typically hot refrigerant vapor, through the evaporator heat exchanger. Defrost may also be accomplished through the activation of one or more electrical resistance heater operatively associated with the evaporator heat exchange tube coil for heating the tube coil.

In operating refrigerant vapor compression systems, knowing when to interrupt a cooling cycle to initiate a defrost cycle is important to operating the refrigerant vapor compression system in a most efficient manner. Initiating a defrost cycle at the expiration of specified time intervals of operation in the cooling mode is a simple, but inefficient, control method. U.S. Pat. No. 6,205,800 discloses a method for defrosting on demand by initiating a defrost routine for removing condensate from an evaporator of a refrigerated device if the difference between the sensed air temperature within the refrigerated enclosure of the refrigerated device and the refrigerant temperature sensed by a refrigerant temperature sensor mounted on or disposed within the evaporator tube coil is greater or equal to a defrost threshold. U.S. Pat. No. 6,318,095 discloses controlling an outdoor coil defrost cycle on a reversible heat pump by continuously monitoring the difference between the outdoor coil temperature and the outdoor air temperature and initiating a defrost cycle when that difference exceeds a target value.

Refrigerant vapor compression systems used in connection with transport refrigeration systems are generally subject to more stringent operating conditions due to the wide range of refrigeration load conditions and the wide range of outdoor ambient conditions over which the refrigerant vapor compression system must operate to maintain product within the cargo space at a desired temperature. The refrigerant vapor compression system must not only have sufficient capacity to rapidly pull down the temperature of product loaded into the cargo space at ambient temperature, but also should operate energy efficiently over the entire load range, including at low load when maintaining a stable product temperature during transport.

The air within the transport cargo box may have a particularly high moisture level after product is first loaded, therefore frost formation can be particularly troublesome during pull down when maximum cooling capacity is needed to draw down the product temperature as quickly as possible. Excessive accumulation of frost on the evaporator tube coil results in reduced heat transfer, which prolongs the time required for pull down. A common method currently in use in truck trailer applications for controlling initiation of a defrost cycle relies on a differential pressure switch that triggers a defrost cycle whenever the airside pressure drop across the evaporator tube coil exceeds a preset threshold.

However, other factors that are not related to frost formation can also impact the airside pressure drop. For example, field installed air chutes can significantly alter the air flow patterns and low airside air flows through the evaporator may not be sufficient to cause the pressure differential switch to trigger despite excessive frost formation of the heat exchange surface of the evaporator tube coil. Further, when the system is operating an low fan speeds, for example such as during a stable temperature maintenance cooling mode or a low noise operational mode, the airside air flow through the evaporator may again be too low to cause the differential pressure switch to rigger despite excessive frost accumulation on the evaporator tube coil.

Additionally, non-uniform frost/ice buildup is not an uncommon problem with respect to evaporators associated with refrigerant vapor compression systems in transport refrigeration applications. As a result of air flow maldistribution through the evaporator heat exchanger, frost/ice buildup may be heavy on some sections of the evaporator heat exchange surface and nearly non-existent on other sections of the evaporator heat exchange surface. The air flow over the heat transfer surface becomes restricted and may not generate enough pressure drop to trigger an air pressure defrost switch to defrost the sections of the evaporator that have heavy frost/ice accumulations. Typically, in transport refrigeration applications, the refrigeration unit is provided with a safety defrost which is automatically triggered whenever the temperature differential between the sensed return air temperature and a sensed evaporator heat exchanger surface temperature exceeds a preselected threshold, which is indicative of insufficient heated being absorbed by the refrigerant due to frost build-up on the evaporator heat exchange surface. The sensed surface temperature is typically taken by a thermister mounted to the heat exchanger tube sheet or a tube fin, but could also be mounted on the surface of a tube.

Continued cooling operation with an excessively frosted coil is inefficient. Cooling capacity may roll-off by 75% or more over as little as two or three hours of operation in the cooling mode with an excessively frosted coil. Continued cooling operation with an excessively frosted coil also results in increased diesel fuel consumption to power the refrigeration unit. Therefore, an active and more direct method for initiating a defrost cycle that is directly influenced by the build-up of frost on the heat exchange surface of the evaporator tube coil is needed.

SUMMARY OF THE INVENTION

A method is provided for controlling initiation of a defrost cycle of an evaporator heat exchanger of a refrigeration system. The method includes the steps of: establishing a return air-saturation temperature differential equal to the difference of a sensed air temperature of an air flow returning from the cargo box to pass over the evaporator heat exchanger minus a refrigerant saturation temperature of a flow of refrigerant passing through the evaporator heat exchanger; comparing the return air-saturation temperature differential to a set point threshold defrost temperature differential; and if the return air-saturation temperature differential exceeds the set point threshold defrost temperature differential, initiating a defrost cycle for defrosting the evaporator heat exchanger.

The method may include the further step of sensing the air temperature of and generating a signal indicative of the sensed air temperature of an air flow returning from the cargo box to pass over the evaporator heat exchanger. In an aspect, the method may further include the steps of: sensing a refrigerant pressure of a flow of refrigerant passing through the evaporator heat exchanger and generating a signal indicative of the sensed refrigerant pressure of the flow or refrigerant passing through the evaporator heat exchanger; and determining the refrigerant saturation temperature based upon the sensed refrigerant pressure signal.

In an aspect, the method may include the further steps of: sensing a refrigerant pressure of and generating a signal indicative of the sensed refrigerant pressure of a flow of refrigerant passing through the evaporator heat exchanger at a plurality of spaced time intervals over a selected time period; calculating a plurality of refrigerant saturation temperatures, one per each one of the plurality of refrigerant pressures sensed over the selected time period; calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures; and establishing the return air-saturation temperature differential as the difference of the sensed air temperature minus the adjusted refrigerant saturation temperature. The step of calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures may include calculating the adjusted refrigerant saturation temperature as an arithmetic average of the plurality of refrigerant saturation temperatures. The step of calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures may include calculating the adjusted refrigerant saturation temperature as an arithmetic mean of the plurality of refrigerant saturation temperatures. In an aspect, the selected time period may range from at least about three minutes up to about five minutes.

In an aspect, the method may include the further step of adjusting the set point threshold defrost temperature differential as a function of refrigerant mass flow rate of the refrigerant flowing through the evaporator heat exchanger prior to comparing the return air-saturation temperature differential to the set point threshold defrost temperature differential. In an aspect, the method may include the further steps of: calculating a clean coil temperature differential equal to the difference of the sensed return air temperature minus the refrigerant saturation temperature following termination of the defrost cycle; resetting the set point threshold defrost temperature to be the clean coil temperature differential plus a predetermined temperature delta; and initiating the next defrost cycle when the return air-saturation temperature differential exceeds the reset set point temperature differential.

BRIEF DESCRIPTION OF THE DRAWINGS

For a further understanding of the disclosure, reference will be made to the following detailed description which is to be read in connection with the accompanying drawing, wherein:

FIG. 1 is a perspective view of a truck trailer equipped with a refrigeration unit operatively associated with a temperature controlled cargo box;

FIG. 2 is a schematic diagram of an exemplary embodiment of a refrigerant vapor compression system associated with the refrigeration unit of the truck trailer of FIG. 1;

FIG. 3 is a schematic illustration of an exemplary embodiment of the evaporator heat exchanger of the refrigerant vapor compression system of FIG. 2; and

FIG. 4 is a block diagram illustrating an exemplary embodiment of the method disclosed herein.

FIG. 5 is a block diagram illustrating an alternate embodiment of the method illustrated in FIG. 4; and

FIG. 6 is a block diagram illustrating an additional step to the method illustrated in FIG. 4.

DETAILED DESCRIPTION OF THE INVENTION

Referring initially to FIG. 1, there is shown a truck trailer 100 having a refrigerated cargo box 110 having access doors 112 that open to the interior space 114 of the cargo box from the exterior of the truck trailer to facilitate loading of product into the cargo box 110 for transport. The truck trailer 100 is equipped with a transport refrigeration unit 10 for regulating and maintaining a temperature controlled atmosphere within the cargo box during transport within a desired storage temperature range selected for the perishable product being shipped therein. Although the demand defrost method disclosed herein will be described herein with reference to the refrigerated cargo box of the depicted truck trailer, it is to be understood that the invention may also be used in connection with other refrigerated cargo transport boxes, including for example a refrigerated box of a truck, or a refrigerated cargo container for transporting perishable product by ship, by rail, by road or intermodal transport. The disclosed demand defrost method may also be applied to controlling evaporator defrost cycle initiation on demand in refrigerant vapor compression systems for supplying conditioned air to a temperature controlled space, such as use in connection with air conditioning systems and commercial refrigeration systems.

Referring now also to FIG. 2, the transport refrigeration unit 10 includes a refrigerant vapor compression system 12 and an associated power source. The refrigerant vapor compression system 12 includes a compression device 20, a condenser 30 having a heat exchanger and associated condenser fan(s) 34, an evaporator 40 having a heat exchanger 42 and associated evaporator fan(s) 44, and an evaporator expansion device 46, all arranged in a conventional refrigeration cycle and connected in a refrigerant circulation circuit including refrigerant lines 22, 24, 26 and the condenser tubular heat exchanger 32 and the evaporator tubular heat exchanger 42. The transport refrigeration system 10 is mounted as in conventional practice to an exterior wall of the truck trailer 100, for example the front wall 116 thereof, with the compressor 20 and the condenser 30 with its associated condenser fan(s) 34 and power source 50 disposed externally of the refrigerated cargo box 110 in a housing 118.

The evaporator 40 extends through an opening in the front wall 116 into the refrigerated cargo box 110. The expansion device 46, which in the depicted embodiment is an electronic expansion valve, but could be a thermostatic expansion valve, is disposed in refrigerant line 24 downstream with respect to refrigerant flow of the condenser heat exchanger 32 and upstream with respect to refrigerant flow of the evaporator heat exchanger 42 for metering the flow of refrigerant through the evaporator in response to the degree of superheat in the refrigerant at the outlet of the evaporator 40, as in conventional practice. A refrigerant pressure sensor 48 is mounted on the tubular heat exchanger 42 of the evaporator 40 for monitoring the sensing the refrigerant flowing through the evaporator heat exchanger 42 at or near the outlet thereof. Although the particular type of evaporator heat exchanger 42 used is not limiting of the invention, the evaporator heat exchanger 42 may, for example, comprise one or more heat exchange tube coils, as depicted in the drawing, or one or more tube banks formed of a plurality of tubes extending between respective inlet and outlet manifolds. The tubes may be round tubes or flat tubes and may be finned or un-finned.

The compressor 20 may comprise a single-stage or multiple-stage compressor such as, for example, a reciprocating compressor or a scroll compressor, although the particular type of compressor used is not germane to or limiting of the invention. In the exemplary embodiment depicted in FIG. 2, the compressor is a reciprocating compressor, such as for example, an 06D model reciprocating compressor manufactured by Carrier Corporation or a variant thereof, having a compressing mechanism, an internal electric compressor motor and an interconnecting drive shaft that are all sealed within a common housing of the compressor 20. The power source 50 powers the internal electric motor of the compressor. In an embodiment, the power source 50 generates sufficient electrical power for fully driving the electrical motor of the compressor 20 and also for providing all other electrical power required by the fans 34, 44 and other parts of the refrigeration unit 10. In an electrically powered embodiment of the transport refrigeration unit 10, the power source 50 comprises a single on-board engine driven synchronous generator configured to selectively produce at least one AC voltage at one or more frequencies. An electrically powered transport refrigeration system suitable for use on truck trailer transport vehicles are shown in U.S. Pat. No. 6,223,546, assigned to the assignee of the present application, the entire disclosure of which is incorporated herein by reference.

The transport refrigeration unit 10 also includes an electronic controller 60 that is configured to operate the transport refrigeration unit 10 to maintain a predetermined thermal environment within the interior space 114 defined within the cargo box 110 wherein the product is stored during transport. The electronic controller 60 maintains the predetermined thermal environment by selectively powering and controlling the operation of various components of the refrigerant vapor compression system, including the compressor 20, the condenser fan(s) 34 associated with the condenser 30, the evaporator fan(s) 44 associated with the evaporator 40, and various valves in the refrigerant circuit, including but not limited to the electronic expansion valve 46 (if present) and the suction modulation valve 62 (if present). When cooling of the environment within interior space 114 of the cargo box 110 is required, the electronic controller 60 activates the compressor 20, the condenser fan(s) 34 and the evaporator fan(s) 44, as appropriate, and adjusts the position of the electronic expansion valve 46 to meter the flow of refrigerant through the evaporator heat exchanger 42 to provide a desired degree of superheat in the refrigerant vapor at the evaporator outlet, and adjusts the position of the suction modulation valve 62 to increase or decrease the flow of refrigerant supplied to the compressor 20 as appropriate to control and stabilize the temperatures within the interior space 114 within the cargo box 110 at the respective set point threshold defrost temperature, which corresponds to the desired product storage temperatures for the particular products stored within cargo box 110.

In one embodiment, the electronic controller 60 includes a microprocessor and an associated memory. The memory of the controller 60 may be programmed to contain preselected operator or owner desired values for various operating parameters within the system, including, but not limited to, a temperature set point for the air within the interior space 114 of the cargo box 110, refrigerant pressure limits, current limits, engine speed limits, and any variety of other desired operating parameters or limits within the system. The programming of the controller is within the ordinary skill in the art. The controller 60 may include a microprocessor board that includes the microprocessor, an associated memory, and an input/output board that contains an analog-to-digital converter which receives temperature inputs and pressure inputs from a plurality of sensors located at various points throughout the refrigerant circuit and the refrigerated cargo box, current inputs, voltage inputs, and humidity levels. The input/output board may also include drive circuits or field effect transistors and relays which receive signals or current from the controller 60 and in turn control various external or peripheral devices associated with the transport refrigeration system. In an embodiment, the controller 60 may comprise the MicroLink™ controller available from Carrier Corporation, the assignee of this application. However, the particular type and design of the controller 60 is within the discretion of one of ordinary skill in the art to select and is not limiting of the invention.

As in conventional practice, when the refrigerant vapor compression system is in operation, low temperature, low pressure refrigerant vapor is compressed by the compressor 20 to a high pressure, high temperature refrigerant vapor and passed from the discharge outlet of the compressor 20 into refrigerant line 22. The refrigerant circulates through the refrigerant circuit via refrigerant line 22 to and through the heat exchange tube coil or tube bank of the condenser heat exchanger 32, wherein the refrigerant vapor condenses to a liquid, and the subcooler 32, thence through refrigerant line 24 through a first refrigerant pass of the refrigerant-to-refrigerant heat exchanger 35, and thence traversing the evaporator expansion device 46 before passing through the evaporator heat exchanger 42 and thence through refrigerant line 26, passing a second refrigerant pass of the refrigerant-to-refrigerant heat exchanger 35 before passing to the suction inlet of the compression device 20.

In flowing through the heat exchange tube coil or tube bank of the evaporator heat exchanger 42, the refrigerant evaporates, and is typically superheated, as it passes in heat exchange relationship the air passing through the airside of the evaporator 40. The air is drawn from within the cargo box 110 by the evaporator fan(s) 44, passed over the external heat transfer surface of the heat exchange tube coil or tube bank of the evaporator heat exchanger 42 and circulated back into the interior space 114 of the cargo box 110. The air drawn from the cargo box 110 is referred to as “return air” and the air circulated back to the cargo box 110 is referred to as “supply air”. It is to be understood that the term “air’ as used herein includes mixtures of air and other gases, such as for example, but not limited to nitrogen or carbon dioxide, sometimes introduced into a refrigerated cargo transport box. A temperature sensor 45 is provided to sense the actual temperature of the return air drawn from the temperature controlled interior space 114 of the cargo box 110 before passing over the evaporator heat exchanger 42.

During operation of the refrigerant vapor compression system in a cooling mode, moisture in the return air will condense onto the heat transfer surface, i.e. surface of the tubes and the fins if finned tubes are present, of the evaporator heat exchanger 42 as the return air is cooled in passing in heat exchange relationship with the refrigerant flowing through the evaporator heat exchanger 42. The condensate will freeze on the heat transfer surface of the evaporator heat exchanger 42 and tend to accumulate as a layer of frost and/or ice on the heat transfer surface of the evaporator heat exchanger 42. As the frost/ice layer builds-up, heat transfer performance of the evaporator heat exchanger 42 deteriorates and the airside flow area through the evaporator heat exchanger 42 becomes more and more restricted. Therefore, operation of the refrigerant vapor compression system in the cooling mode must be interrupted to conduct an evaporator defrost cycle whenever the accumulated frost/ice layer becomes excessive.

Referring now to FIG. 3, an electrical resistance heater 70 is provided in operative association with the evaporator heat exchanger 42 to melt the accumulated frost/ice deposited on the heat transfer surface of the evaporator heat exchanger 42. Whenever a defrost cycle is to be implemented, the controller 60 will deactivate the compression device 20, the condenser fan(s) 34 and the evaporator fan(s) 44 for the duration of the defrost cycle and activate the electrical resistance heater 70 for the duration of the defrost cycle by selectively switching on the supply of electrical power from power source 50.

The controller 60 will terminate the defrost cycle by deactivating, i.e. switching off the supply of electrical power to, the electrical resistance heater 70. The controller 60 may terminate the defrost cycle after a predetermined period of time in operation in the defrost cycle elapses or may terminate the defrost cycle based on a temperature signal from a coil defrost termination sensor indicative of a sensed surface temperature indicative of an external tube surface temperature of the evaporator heat exchanger 42. After termination of the defrost cycle, the controller 60 will return the refrigerant vapor compression system to operation in the cooling mode, by restarting the compression device 20, the condenser fan(s) 34 and the evaporator fan(s) 44. Thus, during defrost cycle operation, not only is the air to the controlled space not being cooled, but the heat transfer surface of the evaporator heat exchanger 42 is also being heated.

Referring now to FIG. 4, in accord with the method disclosed herein, the controller 60 will initiate a defrost cycle based on an return air-saturation temperature differential (RASTD), which is defined as the actual return air temperature (RAT), which is sensed by the return air temperature sensor 45 at step 202, minus the refrigerant saturation temperature (ERST) within the evaporator heat exchanger 42. The controller 60 uses the signal indicative of the sensed return air temperature generated by and received from the return air temperature sensor 45 in controlling operation of the refrigeration unit in the cooling mode and also uses the signal indicative of the sensed evaporator refrigerant pressure (ERP) generated by and received from the pressure sensor 48 at step 204 to calculate the evaporator refrigerant saturation temperature (ERST) for controlling the electronic expansion valve 46 to control superheat. Additionally, in accord with an aspect of the method disclosed herein, the controller 60 will, at step 206, determine the evaporator refrigerant saturation temperature (ERST) based upon the sensed evaporator refrigerant pressure (ERP) sensed by pressure sensor 48 at step 204, and at step 208 calculate the return air-saturation temperature differential (RASTD) by subtracting the evaporator refrigerant saturation temperature (ERST) from the actual return air temperature (RAT) sensed by the return air temperature sensor 45 at step 202.

The controller 60, at step 210, compares the calculated return air-saturation temperature differential (RASTD) to a defrost threshold defrost temperature differential (DTSP). If the calculated return air-saturation temperature differential does not exceed the defrost threshold approach temperature differential at block 212, the controller 60 continues operation of the refrigerant vapor compression system in the refrigeration (cooling) mode and repeats steps 202 through 210. However, if the calculated return air-saturation temperature differential exceeds the defrost threshold defrost temperature differential at block 214, the controller 60, interrupts operation of the refrigerant vapor compression system in the refrigeration (cooling) mode and initiates a defrost cycle to remove frost/ice accumulated on the heat transfer surface of the evaporator heat exchanger 42 in the manner discussed hereinbefore. The controller 60 continues operation of the refrigerant vapor compression system 10 in the defrost cycle until all or at least substantially all of the frost/ice accumulated on the heat transfer surface of the evaporator heat exchanger 42 has been removed.

Referring now to FIG. 5, in an aspect of the method depicted therein, at step 207, the controller 60 may calculate an adjusted evaporator refrigerant saturation temperature as a function of a plurality of instantaneous evaporator refrigeration saturation temperatures (ERSi) sensed at spaced time intervals (step 206) to filter out evaporator superheat control related noise and any influence on control logic. In an embodiment, the controller 60 calculates the adjusted evaporator refrigerant saturation temperature as a running average of a plurality of instantaneous evaporator refrigerant saturation temperatures over a selected period of time. In an embodiment, the controller 60 may calculate the adjusted evaporator refrigerant saturation temperature as an arithmetic mean of a plurality of instantaneous evaporator refrigerant saturation temperatures over a selected period of time. For example, the adjusted evaporator refrigerant saturation temperature may be calculated as the arithmetic average or arithmetic mean of those instantaneous evaporator refrigerant saturation temperatures calculated over the immediately past three to five minutes.

In an aspect of the method disclosed herein, the controller 60 may compensate for variation in refrigerant mass flow rate through the evaporator heat exchanger 42 by adjusting the threshold defrost temperature differential (TDTD) as a function of the refrigerant mass flow rate. For example, the controller 60 may select the threshold defrost return air-saturation temperature differential from an initiation curve of threshold defrost return air-saturation temperature differential versus refrigerant mass flow rate through the evaporator heat exchanger 42. The initiation curve may be empirically developed based on testing of the actual refrigerant vapor compression system in use. In determining whether or not to initiate a defrost cycle, the controller 60 will compare the calculated return air-saturation temperature differential to a adjusted threshold defrost temperature differential selected from the aforementioned initiation curve based on the actual refrigerant mass flow rate through the evaporator heat exchanger 42 associated with the evaporator refrigerant saturation temperature used in calculating the return air-saturation temperature differential. If the calculated return air-saturation temperature differential comprises an adjusted return air-saturation temperature differential based on a plurality of instantaneous return air-saturation temperature differentials, then the evaporator refrigerant mass flow rate associated therewith for purposes of selection of the adjusted threshold defrost temperature differential would be the corresponding average or mean evaporator refrigerant mass flow rate.

In a further aspect of the method disclosed herein, the threshold defrost temperature differential may be selected based on a sensed “clean coil” return air-saturation temperature differential, For example, in implementing this aspect of the method, at the termination of each defrost cycle when the heat exchange surface of the evaporator heat exchanger 42 is substantially frost/ice free, the controller 60 will calculate a “clean coil” return air-saturation temperature differential based upon the then current sensed return air temperature and evaporator refrigerant saturation temperature. The controller 60 would then set the defrost threshold approach temperature differential for triggering the next defrost cycle to be a pre-determined temperature delta from that “clean coil” return air-saturation temperature differential. Thus, to trigger a defrost cycle, the return air-saturation temperature differential would need to exceed the actual “clean coil” return air-saturation temperature differential at termination of the last previous defrost cycle by a pre-determined temperature delta. In this aspect of the method disclosed herein, the initiation of defrost cycles is automatically adapted in response to operating conditions associated with the particular product being shipped, local ambient conditions, loading, air flow variations, and other operational factors that may potentially influence frost/ice formation.

The method for initiating a defrost cycle as discloses relies on information available from conventional sensors that are customarily provided on conventional refrigerant vapor compression systems and therefore does not require the installation of new hardware. Additionally, the method disclosed herein eliminates the need for an air pressure switch for initiating defrost, thereby reducing cost and improving overall reliability. Further, triggering defrost based on return air-saturation temperature differential in accord with the method disclosed herein, allows for more effective and more efficient cooling operation by reducing unnecessary run time in the cooling with a highly frosted evaporator while waiting for a safety type defrost to initiate because the air pressure switch failed to trigger a defrost cycle when needed.

As frost builds up on the tube coil or tube bank of the evaporator heat exchanger 42, the air flow through the evaporator 40 goes down the airside pressure drop increases. Consequently, the refrigerant flowing through the heat exchanger tubes absorbs less heat. Therefore, without as much heat going into the refrigerant, the expansion valve 46 throttles the refrigerant flow passing through the tubes of the evaporator heat exchanger 42 in an attempt to maintain the desired refrigerant superheat, which results in a drop in evaporator refrigerant pressure. Thus, the refrigerant saturation temperature also decreases. As the refrigerant saturation temperature goes lower and lower a s the expansion valve continues to throttle the refrigerant flow, the temperature differential with respect to the sensed return air temperature increases, which will lead to a demand defrost when the threshold defrost temperature differential is exceeded. However, a low refrigerant pressure condition resulting from a low refrigerant flow in the evaporator despite a wide open (for example 90% or more open), which could result from a loss of refrigerant charge, could result in an on demand defrost cycle being called for when the frost build-up per se does not warrant a defrost. Referring now to FIG. 6, to avoid the initiation of an on demand defrost cycle, the controller 60, at step 216, will monitor the position of the expansion valve 46 and the degree of superheat in the refrigerant leaving the evaporator heat exchanger 42 as a feedback to detect whether the one demand defrost indication is the result of a low refrigerant flow condition through the evaporator heat exchanger 42 rather than the result of excessive frost build-up. The controller 60 will determine whether both the position of the expansion valve 46 and the superheat are within normal operating range. If so, the controller 60 will terminate operation in the cooling mode and initiate a defrost cycle. If not, the controller will continue operation in the cooling mode.

The terminology used herein is for the purpose of description, not limitation. Specific structural and functional details disclosed herein are not to be interpreted as limiting, but merely as basis for teaching one skilled in the art to employ the present invention. Those skilled in the art will also recognize the equivalents that may be substituted for elements described with reference to the exemplary embodiments disclosed herein without departing from the scope of the present invention.

While the present invention has been particularly shown and described with reference to the exemplary embodiment as illustrated in the drawing, it will be recognized by those skilled in the art that various modifications may be made without departing from the spirit and scope of the invention. Therefore, it is intended that the present disclosure not be limited to the particular embodiment(s) disclosed as, but that the disclosure will include all embodiments falling within the scope of the appended claims.

Claims

1. A method for controlling initiation of a defrost cycle of an evaporator heat exchanger of a refrigerant vapor compression system for supplying conditioned air to a temperature controlled space, the method comprising the steps of:

establishing a return air-saturation temperature differential equal to the difference of a sensed air temperature of an air flow returning from the temperature controlled space to pass over the evaporator heat exchanger minus a refrigerant saturation temperature of a flow of refrigerant passing through the evaporator heat exchanger;

comparing the return air-saturation temperature differential to a set point threshold defrost temperature differential; and

if the return air-saturation temperature differential exceeds the set point threshold defrost temperature differential, initiating a defrost cycle for defrosting the evaporator heat exchanger.

2. The method as recited in claim 1 further comprising the step of sensing the air temperature of and generating a signal indicative of the sensed air temperature of an air flow returning from the temperature controlled space to pass over the evaporator heat exchanger.

3. The method as recited in claim 1 further comprising the steps of:

sensing a refrigerant pressure of and generating a signal indicative of the sensed refrigerant pressure of a flow of refrigerant passing through the evaporator heat exchanger;

determining the refrigerant saturation temperature based upon the sensed refrigerant pressure signal.

4. The method as recited in claim 1 further comprising the steps of:

sensing a refrigerant pressure of and generating a signal indicative of the sensed refrigerant pressure of a flow of refrigerant passing through the evaporator heat exchanger at a plurality of spaced time intervals over a selected time period;

calculating a plurality of refrigerant saturation temperatures, one per each one of the plurality of refrigerant pressures sensed over the selected time period;

calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures; and

establishing the return air-saturation temperature differential as the difference of the sensed air temperature minus the adjusted refrigerant saturation temperature.

5. The method as recited in claim 4 wherein the step of calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures comprises calculating the adjusted refrigerant saturation temperature as an arithmetic average of the plurality of refrigerant saturation temperatures.

6. The method as recited in claim 4 wherein the step of calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures comprises calculating the adjusted refrigerant saturation temperature as an arithmetic mean of the plurality of refrigerant saturation temperatures.

7. The method as recited in claim 4 wherein the selected time period ranges from at least about three minutes up to about five minutes.

8. The method as recited in claim 1 further comprising the step of adjusting the set point threshold defrost temperature differential as a function of refrigerant mass flow rate of the refrigerant flowing through the evaporator heat exchanger prior to comparing the return air-saturation temperature differential to the set point threshold defrost temperature differential.

9. The method as recited in claim 1 further comprising the steps of:

calculating a clean coil temperature differential equal to the difference of the sensed return air temperature minus the refrigerant saturated temperature following termination of the defrost cycle;

resetting the set point threshold defrost temperature differential to be the clean coil temperature differential plus a predetermined temperature delta; and

initiating the next defrost cycle when the return air-saturation temperature differential exceeds the reset set point threshold defrost temperature differential.

10. The method as recited in claim 1 further comprising the step of determining that the position of an evaporator expansion valve is within normal operating range prior to initiating a demand defrost.

11. A method for controlling initiation of a defrost cycle of an evaporator heat exchanger of a refrigeration system operatively associated with a refrigerated transport cargo box, the method comprising the steps of:

establishing a return air-saturation temperature differential equal to the difference of a sensed air temperature of an air flow returning from the cargo box to pass over the evaporator heat exchanger minus a refrigerant saturation temperature of a flow of refrigerant passing through the evaporator heat exchanger;

comparing the return air-saturation temperature differential to a set point threshold defrost temperature differential; and

if the return air-saturation temperature differential exceeds the set point threshold defrost temperature differential, initiating a defrost cycle for defrosting the evaporator heat exchanger.

12. The method as recited in claim 11 further comprising the step of sensing the air temperature of and generating a signal indicative of the sensed air temperature of an air flow returning from the cargo box to pass over the evaporator heat exchanger.

13. The method as recited in claim 11 further comprising the steps of:

sensing a refrigerant pressure of and generating a signal indicative of the sensed refrigerant pressure of a flow of refrigerant passing through the evaporator heat exchanger;

determining the refrigerant saturation temperature based upon the sensed refrigerant pressure signal.

14. The method as recited in claim 11 further comprising the steps of:

sensing a refrigerant pressure of and generating a signal indicative of the sensed refrigerant pressure of a flow of refrigerant passing through the evaporator heat exchanger at a plurality of spaced time intervals over a selected time period;

calculating a plurality of refrigerant saturation temperatures, one per each one of the plurality of refrigerant pressures sensed over the selected time period.

calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures; and

establishing the return air-saturation temperature differential as the difference of the sensed air temperature minus the adjusted refrigerant saturation temperature.

15. The method as recited in claim 14 wherein the step of calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures comprises calculating the adjusted refrigerant saturation temperature as an arithmetic average of the plurality of refrigerant saturation temperatures.

16. The method as recited in claim 14 wherein the step of calculating an adjusted refrigerant saturation temperature based on the plurality of refrigerant saturation temperatures comprises calculating the adjusted refrigerant saturation temperature as an arithmetic mean of the plurality of refrigerant saturation temperatures.

17. The method as recited in claim 14 wherein the selected time period ranges from at least about three minutes up to about five minutes.

18. The method as recited in claim 11 further comprising the step of adjusting the set point threshold defrost temperature differential as a function of refrigerant mass flow rate of the refrigerant flowing through the evaporator heat exchanger prior to comparing the return air-saturation temperature differential to the set point temperature differential.

19. The method as recited in claim 11 further comprising the steps of:

calculating a clean coil temperature differential equal to the difference of the sensed return air temperature minus the refrigerant saturated temperature following termination of the defrost cycle;

resetting the set point threshold defrost temperature to be the clean coil temperature differential plus a predetermined temperature delta; and

initiating the next defrost cycle when the return air-saturation temperature differential exceeds the reset set point threshold defrost temperature differential.

20. The method as recited in claim 11 further comprising the step of determining that the position of an evaporator expansion valve is within normal operating range prior to initiating a demand defrost.