ENHANCED REFRIGERANT SYSTEM

- CARRIER CORPORATION

A refrigerant system providing enhanced performance over a wider range of operating conditions than traditional economized refrigerant systems. The system includes an economizer branch that connects a liquid outlet from a suction accumulator to an economizer inlet port of a compressor unit. The economizer branch includes a liquid refrigerant pump that delivers a non-evaporated liquid refrigerant portion from the suction accumulator into the economizer heat exchanger, where the liquid refrigerant portion evaporates, increasing thermodynamic potential of the main circuit refrigerant also flowing in through the economizer heat exchanger, and a formed vapor stream is delivered into the economizer inlet port of the compressor unit.

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Description
FIELD OF THE INVENTION

The present invention relates to refrigerant systems, and particularly to refrigerant systems employing economizer cycles.

BACKGROUND OF THE INVENTION

Refrigerant systems are employed to control conditions, such as temperature and humidity, in targeted spaces. Some refrigerant systems are configured as heat pumps to perform heating or cooling duties on demand. The performance (capacity and/or efficiency) of a refrigerant system can be enhanced by employing an economizer cycle. (See U.S. Pat. No. 6,385,981B1, U.S. Pat. No. 6,571,576B1, and U.S. Pat. No. 7,000,423B2).

SUMMARY OF THE INVENTION

In one aspect, the invention provides an enhanced refrigerant system that includes a refrigerant closed-loop circuit comprising a compressor unit, a heat rejection unit, an economizer heat exchanger, an expansion device, an evaporator unit, and a suction accumulator. The suction accumulator includes an inlet, a vapor outlet, and a liquid outlet. The compressor unit includes a suction inlet port, an economizer inlet port, and a discharge outlet port.

An economizer line provides a passage of a refrigerant stream flowing between the liquid outlet of the suction accumulator and the economizer inlet port of the compressor unit. The economizer line includes a liquid refrigerant pump and an economizer heat exchanger.

The economizer heat exchanger provides heat transfer interaction between the refrigerant stream flowing in the economizer line and the refrigerant stream flowing in the main refrigerant loop. The evaporator unit is configured and operated in such way that at least a portion of refrigerant leaving the evaporator unit is in a liquid phase.

The liquid refrigerant pump of the economizer line pumps the liquid refrigerant, that is at least a portion of refrigerant flow exiting the evaporator unit, through the economizer line and the economizer heat exchanger. At least a portion of this liquid refrigerant evaporates in the economizer heat exchanger and forms a vapor refrigerant stream in the economizer line that flows into the economizer inlet port of the compressor unit.

The evaporator unit is configured and operated to provide at least some non-evaporated liquid refrigerant portion at the evaporator unit outlet. The liquid refrigerant pump delivers this non-evaporated liquid refrigerant portion into the economizer heat exchanger, where the liquid refrigerant portion is at least partially evaporated and delivered into the economizer inlet port of the compressor unit.

If the enhanced refrigerant system is charged and operated with a sub-critical fluid, the heat rejection unit is called a condenser. If the enhanced refrigerant system is charged and operated with a trans-critical refrigerant, the heat rejection unit is a so-called gas cooler.

The compressor unit, the heat rejection unit, the expansion device unit, the evaporator unit, the economizer heat exchanger unit, the suction accumulator unit, and/or liquid refrigerant pump unit may have pluralities of components within these units such as compressors, heat rejection heat exchangers, expansion devices, evaporators, suction accumulators, and liquid refrigerant pumps respectively.

If the enhanced refrigerant system is used for cooling, then the heat rejection unit is an outdoor unit and the evaporator unit is an indoor unit. If the enhanced refrigerant system is used for heating, then the heat rejection unit is an indoor unit and the evaporator unit is an outdoor unit.

If the refrigerant system is used as a heat pump, that is, namely for heating and cooling, a four-way reversing valve is employed to redirect refrigerant flow while switching between cooling and heating modes of operation. The four-way reversing valve has a vapor inlet, a vapor outlet, a first bi-flow port, and a second bi-flow port. The vapor inlet is connected to the discharge port of the compressor unit. The vapor outlet is connected to the vapor outlet of the suction accumulator. The first bi-flow port is connected to the outdoor unit, and the second bi-flow port is connected to the indoor unit.

Some refrigerant system may be combined into a single unit. For instance, the compressor unit and the heat rejection unit may be assembled as a one unit. Also, the expansion device may be combined with the evaporator unit. Further, the liquid refrigerant pump and the suction accumulator may be combined as well.

The compressor unit may have two compressors, namely a low-pressure compressor and a high-pressure compressor, with the economizer inlet port positioned between these compressors. Each of the compressors has at least one compression stage. Each compression stage may have a plurality of parallel so-called tandem compressors. The low-pressure compressor and the high-pressure compressor may be manufactured and assembled as separate units or as a single unit.

The economizer heat exchanger may have a counterflow, a parallel flow, or a crossflow arrangement. It may be replaced by a flash tank as well. The flash tank has a vapor inlet port, a vapor outlet port, and two liquid ports. It provides a direct thermal contact between a refrigerant stream flowing in the main refrigerant loop and a refrigerant stream flowing in the economizing line. At least one of two liquid ports has an expansion device located within the main refrigerant loop upstream to the flash tank. At least one of two liquid ports and the flash tank comprise a single unit.

If the refrigerant system employs a trans-critical refrigerant, the economizer inlet port of the compressor unit may be combined with the discharge port of the compressor unit. In this case, the liquid refrigerant pump will operate in parallel to the compressor unit.

The enhanced refrigerant system has the following advantages in comparison with the traditional economized systems: 1) a portion of the total refrigerant mass flow rate is pumped by the liquid pump, which requires substantially lower power input; 2) the total refrigerant mass flow circulating throughout the refrigerant system is delivered through the evaporator unit increasing the evaporator capacity; 3) on the contrary to the conventional systems, the higher the economizing pressure is, the better capacities, compressor power, and coefficient of performance (COPs) are; 4) when ambient temperature is reduced, density of liquid refrigerant at the pump inlet is increased that, along with the pumping capacity, counterbalances degradation of heating capacity and COP; and 5) when ambient temperature is elevated, the economizing pressure is increased diminishing the extent of degradation of cooling capacity and COP.

As a result, the enhanced refrigerant system provides enhance heating and cooling capacities and heating and cooling COPs over a wider range of operating conditions than the traditional economized systems.

This system design yields an enhanced heating and cooling coefficient of performance (COP) and a higher system capacity than the traditional economized refrigerant systems. In particular, the proposed enhanced refrigerant system provides performance enhancements at low ambient temperatures, in a heating mode of operation, and at high temperatures, in a cooling mode of operation.

In the drawings as hereinafter described, an embodiment is depicted; however various other modifications and alternate constructions can be made thereto without departing from the true spirit and scope of the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention can be better understood with reference to the claims and drawings described below. The drawings are not necessarily to scale, the emphasis is instead generally being placed upon illustrating the principles of the invention. Within the drawings, like reference numbers are used to indicate like parts throughout the various views. Differences between like parts may cause those parts to be indicated by different reference numbers. Unlike parts are indicated by different reference numbers.

FIG. 1 illustrates a traditional (prior art) refrigerant system including an economizer cycle;

FIG. 2 illustrates a traditional (prior art) heat pump including an economizer cycle;

FIG. 3 illustrates a traditional (prior art) refrigerant system;

FIG. 4 illustrates an enhanced refrigerant system in accordance with the invention;

FIG. 5 illustrates a pressure-enthalpy diagram of the enhanced refrigerant system;

FIG. 6 illustrates a boiling point elevation demonstration;

FIG. 7A illustrates an enhanced refrigerant system including sequential multi-stage compression;

FIG. 7B illustrates an enhanced refrigerant system including parallel multi-stage compression;

FIG. 8A illustrates an enhanced refrigerant system operating with a trans-critical refrigerant;

FIG. 8B illustrates an enhanced refrigerant system including a split economizer port;

FIG. 8C illustrates an enhanced refrigerant system including a liquid receiver located downstream a condenser, with respect to a refrigerant flow;

FIG. 8D illustrates an enhanced refrigerant system including a liquid receiver located between a condenser and a sub-cooler;

FIG. 9 illustrates an enhanced heat pump;

FIG. 10 illustrates the enhanced heat pump with a reversing flow arrangement through the economizer heat exchanger.

FIG. 11 illustrates the enhanced heat pump with a flash tank.

 DETAILED DESCRIPTION OF THE INVENTION

A refrigerant system typically comprises a refrigerant closed-loop circuit with the following components connected in sequence: a compressor, a heat rejection unit, an expansion device, an evaporator unit, and quite frequently a suction accumulator. The evaporator unit provides heat transfer interaction between an evaporating refrigerant stream at low pressure and temperature and a secondary fluid being delivered to a climate controlled space from which heat is to be rejected and transferred to this evaporating refrigerant stream. The heat rejection unit provides a heat transfer interaction between a compressed refrigerant stream at high pressure and temperature and another secondary fluid flowing in the environment to which heat is to be rejected from this compressed refrigerant stream.

When traditional (sub-critical) refrigerants are utilized in refrigerant systems, the heat rejection unit is referred to as a condenser. In a condenser, at least a portion of the heated and compressed refrigerant stream is liquefied from a vapor phase. When trans-critical refrigerants are employed, the heat rejection unit is referred to as a gas cooler (See International Patents WO9007683 and WO9306423). In a gas cooler, the compressed refrigerant at high pressure and temperature is maintained in a gaseous phase.

Suction accumulators are often incorporated into refrigerant systems when flow rates of supplying to the evaporating unit liquid refrigerant may result in an imbalance with the thermal load. The imbalance can cause liquid refrigerant to be present at the evaporator outlet and the compressor inlet, which may compromise compressor reliability. Also, the suction accumulators are applied when a substantial amount of oil-refrigerant mixture can intermittently accumulate at the evaporator outlet.

Heat pumps are intended to enable heating or cooling duties, also referred to as heating or cooling modes of operation. A heat pump consists of a refrigerant closed-loop circuit with the following components connected in sequence: a compressor, a four-way reversing valve, an outdoor heat exchanging unit, an expansion device, an indoor heat exchanging unit, and a suction accumulator.

In the cooling mode, the four-way reversing valve directs a refrigerant stream in such a way that the outdoor heat exchanger unit operates as a heat rejection unit (implemented as a condenser in sub-critical applications or a gas cooler in trans-critical applications). The indoor heat exchanger unit operates as a heat accepting unit, implemented as an evaporator, providing the cooling duty.

In the heating mode, the outdoor heat exchanger unit operates as a heat accepting unit, implemented as an evaporator. The indoor unit operates as a heat rejection unit (once again, implemented as a condenser in sub-critical applications or a gas cooler in trans-critical applications), providing the heating duty.

The heating capacity and a coefficient of performance (COP) of a heat pump are reduced when the ambient temperature is decreased due to following reasons. When the ambient temperature is decreased, the suction pressure and the vapor refrigerant density at the compressor suction are reduced in response to a vapor refrigerant pressure reduction caused by the lower ambient temperature. However, the pressure ratio (discharge pressure divided by the suction pressure) across the compressor is increased. The increased pressure ratio may result in the reduced volumetric efficiency of the compressor, and this fact, together with the reduced refrigerant density, cause a refrigerant mass flow rate reduction and a capacity degradation of the entire system. Also, the increased pressure ratio causes the compressor to work harder and to consume more power in order to pump a unit of refrigerant mass, causing COP degradation.

The cooling capacity and COP of a heat pump are reduced when the ambient temperature is elevated, since the discharge pressure of the compressor and pressure ratio associated with the compressor are increased, causing the compressor to work harder and to consume more power. Elevated ambient temperatures also affect operation of heat rejection heat exchanger, consequently reducing cooling thermal potential of refrigerant entering an evaporator.

In order to limit the extent of the heating capacity and COP reduction, the heat pumps employ an economizer cycle (See U.S. Pat. No. 6,385,981B1, U.S. Pat. No. 6,571,576B1, and U.S. Pat. No. 7,000,423B2). A heat pump with the economizer cycle typically comprises a compressor with a suction inlet port and an economizer inlet port, and consists of a refrigerant closed-loop circuit with the following components connected in a sequence: the compressor, the four-way reversing valve, the outdoor heat exchanger unit, an economizer heat exchanger, the expansion device, the indoor heat exchanger unit, and the suction accumulator. The heat pump typically has a refrigerant circuit branch connecting an outlet of the heat rejection unit and the economizer inlet port of the compressor.

This economizer branch includes an economizer expansion device and the economizer heat exchanger. The economizer heat exchanger provides heat transfer interaction between a refrigerant stream flowing from the heat rejection unit and an evaporating refrigerant stream expanded to some intermediate (between suction and discharge) pressure and temperature flowing from the economizer expansion device. This arrangement increases the mass flow rate through the heat rejection unit and enhances (increases) the heating capacity of the heat pump, in the heating mode of operation. Also, the compressor power is increased as well, and an adequate heating COP may be maintained within some (although still somewhat limited) range of operating conditions.

In the cooling mode, the economizer cycle increases the cooling capacity, but the power required to operate the compressor is increased as well, and as a result, the cooling COP typically doesn't change appreciably (unless special provisions or design features are incorporated).

It has to be pointed out that various economized heat pump designs are feasible and known in the art, which provide similar benefits. Such design schematics may include heat pumps with the economizer expansion devices positioned upstream or downstream of the economizer heat exchanger with respect to refrigerant flow, heat pumps with dual economizer heat exchangers, heat pumps with dual four-way reversing valves, etc. All these schematics are within the scope and can equally benefit from the present invention.

The present invention enhances heating and cooling capacities and heating and cooling COPs of refrigerant systems like those described above and provides operation of an enhanced system over a wider range of operating conditions than those of the traditional economized systems.

FIG. 1 illustrates a traditional (prior art) refrigerant system 100 incorporating an economizer cycle. A traditional refrigerant system with an economizer cycle (FIG. 1) employs a compressor unit 101 having an economizer inlet port 101a. The refrigerant system consists of a refrigerant closed-loop circuit with following components connected in sequence: a compressor unit 101, a heat rejection unit 102, an economizer heat exchanger 103, an expansion device 104, an evaporator unit 105, and a suction accumulator 106. Also, the refrigerant system has an economizer branch 107 fluidly connected to and positioned downstream of an outlet of the heat rejection unit 102. The economizer branch 107 includes an economizer expansion device 108 and the economizer heat exchanger 103 and leads to the economizing inlet port 101a of the compressor unit 101.

If a sub-critical refrigerant is utilized in the refrigerant system 100, the heat rejection unit 102 is a condenser. The economizer heat exchanger 103 provides heat transfer interaction between a liquid refrigerant stream flowing in a refrigerant conduit 103a and an evaporating refrigerant stream flowing in a refrigerant conduit 103b. The heat transfer interaction generates a sub-cooled refrigerant with a reduced enthalpy at an inlet to the expansion device 104 and increased cooling thermal potential in an evaporator 105.

If a trans-critical refrigerant is utilized in the refrigerant system 100, the heat rejection unit 102 is a gas cooler. It provides a high-pressure vapor at a temperature higher than, but close to an ambient or a cooling fluid temperature, at its exit. In this case, the economizer heat exchanger 103 provides heat transfer interaction between a vapor refrigerant stream in the refrigerant conduit 103a and typically an evaporating refrigerant stream in the refrigerant conduit 103b. The heat transfer interaction provides additional cooling of the vapor refrigerant stream in the channel 103a. It has to be pointed out that refrigerant thermodynamic state in the refrigerant conduit 103b, after expansion in the economizer expansion device 108, may be super-critical. In this case, this refrigerant is simply heated during heat transfer interaction in the economizer heat exchanger 103, rather than being evaporated, as explained above.

If the refrigerant system used for heating, the heat rejection unit 102 supplies heat into the climate controlled environment, and the refrigerant system efficiency in a heating mode of operation is evaluated as a ratio of the heat rejection capacity to the total power input. If the refrigerant system used for cooling, the evaporating unit 105 provides cooling (and quite often dehumidification) into the conditioned environment, and the system efficiency in a cooling mode of operation is evaluated by a cooling COP as a ratio of the cooling capacity to the total power input. The total power is comprised of power inputs for all operating electrical components, such as, compressor(s), fan(s), blower(s), and pump(s), etc.

Let us assume that the refrigerant mass flow rate through the evaporator unit 105 is Go. Then the compressor unit 101 receives the refrigerant flow Go through its suction port and discharges the increased (due to an economizer refrigerant flow component Ge) refrigerant flow equal to (Go+Ge). The condenser capacity and compressor power are increased accordingly. Usually the portion Ge is relatively small, and the overall impact of these increases on the heating COP is adequate.

The refrigerant stream at the outlet of the refrigerant conduit 103a, and at the inlet to the expansion device 104, has a lower enthalpy and, therefore, the economizer heat exchanger 103 increases the evaporator capacity. However, the compressor power is also increased to some extent, and the overall effect on the cooling COP may be inadequate.

The lower the economizer pressure is, the higher the economizer heat exchanger, evaporator, and condenser capacities are. On other hand, the compressor power is higher as well. Also, it is true that the higher the economizer pressure is, the lower the economizer heat exchanger, evaporator, and condenser capacities and the compressor power are. Since the capacities and the power are contradicting factors in the COP equation, an appropriate optimal economizer pressure may be selected based on a tradeoff and sensitivity analysis.

FIG. 2 illustrates a traditional (prior art) heat pump 200 including an economizing cycle. The heat pump consists of a refrigerant closed-loop circuit with following components connected in sequence: the compressor unit 101 with the economizer inlet port 101a, a four-way valve 209, an indoor heat exchanger unit 210 operating as a condenser or as a gas cooler, an expansion device assembly 211, the economizer heat exchanger 103, an expansion device assembly 212, an outdoor heat exchanger unit 213 operating as an evaporator unit, and the suction accumulator 106. The heat pump has the economizer branch 107 leading from an outlet of the indoor unit 210, through the economizer expansion device 108, the economizer heat exchanger 103, to the economizer inlet port 101a of the compressor unit 101.

The four-way reversing valve 209 has an inlet port 209a, an outlet port 209b, and two bi-flow ports 209c and 209d.

In the heating mode of operation, the by-flow port 209c is an inlet and the bi-flow port 209d is an outlet. The four-way valve 209 receives the refrigerant vapor stream from the outdoor heat exchanger unit 213 (which operates as an evaporator) via the bi-flow port 209c and directs it to the suction accumulator 106 via the outlet port 209b. The compressor unit 101 receives this refrigerant vapor stream from the suction accumulator 106, compresses and discharges it via the inlet port 209a. The four-way reversing valve 209 directs the received compressed vapor to the indoor heat exchanger unit 210 (which, once again, operates as a condenser or as a gas cooler) via the bi-flow port 209d. In this case, the expansion device assembly 211 is not actuated; the expansion device of the expansion device assembly 212 expands the refrigerant stream to lower pressure and temperature.

In the cooling mode of operation, the bi-flow port 209c is an outlet and the bi-flow port 209d is an inlet. The four-way reversing valve 209 receives the refrigerant vapor stream from the indoor heat exchanger unit 210 (which now operates as an evaporator) via the bi-flow port 209d and directs it to the suction accumulator 106 via the outlet port 209b. Again, the compressor unit 101 receives this refrigerant vapor stream from the suction accumulator 106, compresses and discharges it via the inlet 209a. The four-way reversing valve 209 directs the received compressed vapor to the outdoor heat exchanger unit 213 (which, in the cooling mode, operates as a condenser or a gas cooler) via the bi-flow port 209c. In this case, the expansion device assembly 212 is not actuated; the expansion device of the expansion device assembly 211 expands the refrigerant stream to lower pressure and temperature. As known in the art, expansion device assemblies 211 and 212 comprise an expansion device and a bypass line around this expansion device, with a check valve positioned on the bypass line and allowing refrigerant flow only in one direction.

As stated above the FIG. 2 heat pump schematic is exemplary, and many variations and design options are feasible and within the scope of the invention. These options may include (but are not limited to) heat pumps with the economizer expansion devices positioned upstream or downstream of the economizer heat exchanger 103 with respect to refrigerant flow, heat pumps with dual economizer heat exchangers, heat pumps with dual four-way reversing valves, etc. All these schematics can equally benefit from the present invention.

FIG. 3 illustrates a traditional (prior art) refrigerant system 100 in accordance with the invention and FIG. 4 illustrates an enhanced refrigerant system 400 in accordance with the invention. Referring to FIGS. 3-4, in accordance with the invention, the enhanced refrigerant system consists of the compressor unit 101, the heat rejection unit 102, the refrigerant conduit 103a of the economizing heat exchanger 103, the expansion device 104, the evaporator unit 105, and the suction accumulator 106. The economizer branch 107 connects a liquid outlet refrigerant line 315 from the suction accumulator 106 to the economizer inlet port 101a of the compressor unit 101. The economizer branch 107 includes a liquid refrigerant pump 314 and the refrigerant conduit 103b of the economizer heat exchanger 103.

The compressor unit 101, the heat rejection unit 102, the expansion device 104, the evaporator unit 105, the suction accumulator 106, and liquid refrigerant pump 314 may include associated pluralities of components such as compressors, heat rejection units, expansion devices, evaporators, suction accumulators, and/or liquid refrigerant pumps respectively.

The enhanced refrigeration system may have different design options and enhancement features.

The compressor 101 may be of an open drive, a semi-hermetic, or a hermetic type. It may also employ various compression techniques and include an oil separator, discharge and/or temperature switches. Moreover, the compressor unit 101 may be combined with the heat rejection unit 102.

The heat rejection unit 102 may be cooled by air or by any other secondary fluid. The evaporator unit may cool air or any other secondary fluid as well. The heat transfer to secondary fluid may be of free or forced convection type. The forced convection may be provided by fan(s), blower(s), or pump(s). An expansion device 104 may be a part of the evaporator unit 105.

Each of the refrigerant conduits 103a and channel 103b may include a plurality of channels. The flow arrangement in the heat exchanger may be of a counterflow, parallel flow, or crossflow type, and is defined by a particular application.

The liquid refrigerant pump 314 may be combined into a single unit with the suction accumulator. The pump itself may be of an open drive, a semi-hermetic, or a hermetic (canned) type and may employ various pumping techniques. Further, it may be located inside or outside the suction accumulator. It is feasible to attach the pump to the bottom, the top, or a side wall of the suction accumulator, whether inside or outside the suction accumulator.

The compressor unit 101 receives a refrigerant vapor stream from the suction accumulator 106 at a suction pressure and a refrigerant vapor stream from the economizer branch 107 at the economizer pressure, which is higher then suction pressure, via the economizer inlet port 101a. The hot compressed vapor stream at a high discharge pressure is delivered to the heat rejection unit 102.

If a sub-critical refrigerant is utilized in the refrigerant system 400, the heat rejection unit 102 is a condenser, and the hot compressed vapor refrigerant stream is, at least partially, liquefied. On the other hand, if a trans-critical refrigerant is used, the heat rejection unit 102 is a gas cooler, and the hot compressed refrigerant vapor stream is cooled to a temperature, which is close to and higher than an ambient temperature or a temperature of a cooling fluid.

Additional cooling of the refrigerant stream at the discharge pressure is provided in the refrigerant conduit 103a of the economizer heat exchanger 103 due to the evaporation (or/and heating) of the liquid refrigerant stream in the refrigerant conduit 103b pumped by the liquid refrigerant pump 314 from the suction accumulator 106. This refrigerant flowing through the refrigerant conduit 103b is at lower temperature and pressure than the refrigerant flowing through the refrigerant conduit 103a.

After the expansion from the discharge pressure to the suction pressure in the expansion device 104, the liquid portion of the formed two-phase refrigerant stream is evaporated in the evaporator unit 105. The evaporator unit 105 is sized, configured and operated in such a way that the liquid portion of the refrigerant does not evaporate completely, while flowing through the evaporator unit. The non-evaporated portion is delivered into the suction accumulator 106 and is pumped by the pump 314 through the economizer heat exchanger 103. In the refrigerant conduit 103b of the economizer heat exchanger 103, the liquid stream is evaporated (or/and heated) accepting heat from the refrigerant stream flowing through the refrigerant conduit 103a. The evaporated (or/and heated) refrigerant is received by the economizer port 101a of the compressor unit 101.

A table that compares refrigerant mass flow rates in the traditional and enhanced refrigerant systems is shown below. The table has the following designations: G1 represents a mass flow rate at a location of the suction inlet port to the compressor unit 101, G2 represents a mass flow rate at a location of the compressor discharge, G3 represents a mass flow rate at an exit location of the evaporator unit, and G4 represents a mass flow rate at a location of the economizer inlet port 101a of the compressor unit 101.

G0 and Ge each represent a separate mass flow rate value that is not necessarily associated with a particular location within either the traditional or enhanced systems. G0 represents a mass flow rate value for the evaporator line. Ge represents a mass flow rate for the economizing line.

Based upon the design of each respective refrigerant system, the mass flow rate value for the evaporator line (G0) is equal to the mass flow rate value at the exit location of the evaporator unit of each respective system.

For the traditional system, the mass flow rate value at the exit location of the evaporator unit and at the suction inlet port of the compressor are equal. This is not true for the enhanced system. For the enhanced system, the mass flow rate value at the suction inlet port of the compressor is less than the mass flow rate at the exit location of the evaporator unit.

The mass flow rate value at the location of the discharge of the compressor for the traditional system, represented as (G2 (traditional)=G0+Ge), and for the enhanced system, represented as (G2 (enhanced)=G0), are each equal to the maximum mass flow rate value for each system.

Equivalent Mass Flow Rate Through Case G1 G2 G3 G4 Evaporator Traditional G0 G0 + Ge G0 Ge Enhanced G0 − Ge G0 G0 Ge Compressor Traditional 1 (G0 + Ge)/G0 1 Ge/ G0 Enhanced 1 G0/(G0 − Ge) G0/(G0 − Ge) Ge/(G0 − Ge) Condenser Traditional G0/(G0 + Ge) 1 G0/(G0 + Ge) Ge/(G0 + Ge) Enhanced (G0 − Ge)/G0 1 1 Ge/G0 G0 is mass flow rate through an evaporator in a traditional cycle G0 is mass flow rate through an economizer port in a traditional cycle

Referring to the (2) “Evaporator” associated rows of the table, if we compare refrigerant systems which have the same mass flow rates at the exit location of the evaporator unit (G3 (traditional)=G3 (enhanced)=G0) then the mass flow rate pumped by the compressor unit of the enhanced system, as measured by a mass flow rate at a location of its suction inlet port ((G1 (enhanced)), appears to be smaller than the mass flow rate at the suction inlet port of the compressor for the traditional system (G1 (traditional)).

Specifically, for the enhanced system, at the location of the suction inlet port of the compressor, the mass flow rate is represented by (G1 (enhanced)), which is equal to (G0−Ge). For the traditional system, the mass flow rate at the suction inlet port of the compressor is equal to (G1 (traditional)=G0). Hence, the compressor of the enhanced system pumps a mass flow rate (G1=(G0−Ge)) through its suction inlet port that is less than the mass flow rate (G0) pumped through the suction inlet port by the compressor of the traditional system.

For the enhanced system, the mass flow rate (G1=(G0−Ge)) pumped through the suction inlet port of the compressor is less than the mass flow rate (G0) pumped through the evaporator outlet of the enhanced system (G3 (enhanced)=G0). This is not true for the traditional system. For the traditional system, the mass flow rate pumped through the suction inlet port of the compressor (G1 (traditional) is equal to the mass flow rate pumped through the evaporator exit (G3 (traditional)=G0) of the traditional system.

For the enhanced system, at the location of the discharge port of the compressor, the mass flow rate is represented by (G2 (enhanced)), which is equal to (G1 (enhanced)+Ge) and equal to (G0), which is less than (G2 (traditional)=(G0+Ge)) for the discharge port of the compressor of traditional system. Hence, the compressor of the enhanced system pumps a lower mass flow rate through its discharge port than the mass flow rate pumped through the discharge port by the compressor of the traditional system.

The equivalence of mass flow rates in the evaporating units is associated with the equivalence of the evaporator capacities for the traditional and enhanced systems. The reduced mass flow rate pumped through the compressor unit of the enhanced system indicates that a reduced amount of compressor power is required for the enhanced system. Also, pumping of liquid refrigerant by a liquid pump requires substantially less power than compression of an equivalent mass of vapor and, as a result, the enhanced system yields an increased cooling coefficient of performance (COP).

Referring to the (2) compressor associated rows of the above table, for the enhanced system, an increase in Ge value yields an increase in cooling capacity. An advantage of the enhanced refrigerant system is an opportunity to increase Ge and improve the system performance to a greater extent than the traditional economized systems will allow.

If we compare systems which have the same mass flow rates at a location of the suction inlet port of the compressor unit (G1 (traditional)=G1 (enhanced)=1.0) then the mass flow rate pumped by the compressor unit of the enhanced system, as measured by a mass flow rate at a location of its discharge outlet port ((G2 (enhanced)), appears larger than the mass flow rate at the discharge outlet port of the compressor for the traditional system (G2 (traditional)). Also, at the exit location of the evaporator unit G3 (enhanced) appears to be larger than the mass flow rate at evaporator unit exit for the traditional unit G3 (traditional).

Specifically, for the enhanced system, at the location of the discharge outlet port of the compressor, the mass flow rate is represented by (G2 (enhanced)), which is equal to G0/(G0−Ge). For the traditional system, the mass flow rate at the discharge outlet port of the compressor is equal to (G2 (traditional)=(G0+Ge)/G0<G0/(G0−Ge)). The reduction of mass flow rates at the evaporator unit exit shows that G3 (enhanced)=G0/(G0−Ge)<1=G3 (traditional). Hence, the heat rejection unit 102 and the evaporator unit 105 of the enhanced system handle higher capacities than the same units of the traditional system.

The higher the economizer pressure in the enhanced system is, the better performance characteristics of the enhanced system are, which is completely opposite for the traditional system. It means that the enhanced system may require less power to pump refrigerant from the economizer inlet port to the discharge outlet port, in comparison to the traditional system. This indicates that the enhanced system will also have a COP advantage in cooling and heating modes of operation.

The equivalence of mass flow rates at the compressor suction indicates that application of equivalent compressors in the enhanced and traditional economized systems would be adequate to support the operation of each system. Another advantage of the enhanced refrigerant system is that higher economizer pressure yields higher system capacity and compressor power. As stated above, this is not true for the traditional system.

Referring to the (2) condenser associated rows of the above table, comparing refrigerant systems which have the same mass flow rates at a location of the discharge of the compressor unit (G2 (traditional)=G2 (enhanced)=1.0), which indicates equivalence of condenser heat rejection capacity, the mass flow rate pumped by the compressor unit of the enhanced system, as measured by a mass flow rate at the location of its suction inlet port ((G1 (enhanced)), appears to be smaller than the mass flow rate at the suction inlet port of the compressor for the traditional system (G1 (traditional)). At the exit location of the evaporator unit G3 (enhanced) appears to be larger than the mass flow rate at the evaporator unit exit for the traditional unit G3 (traditional).

The above described advantages provided by the enhanced system allows for having a lower discharge-to-economizer pressure ratio that may be associated with an increased refrigerant mass flow rate. As a result, the enhanced system has a lower required compressor power and an increased heating COP for the same economizer mass flow rate (G4 (enhanced)=G4 (traditional)=Ge). The enhanced system design provides flexibility to incorporate a larger economizer mass flow rate (Ge), in order to enhance the performance characteristics of the enhanced refrigerant system to a greater extent than previously described.

FIG. 5 illustrates a pressure-enthalpy diagram 500 of the enhanced refrigerant system. The pressure-enthalpy diagram 500 demonstrates the following thermodynamic processes, with respect to saturation lines separating sub-critical and trans-critical refrigerants: 501-502 is the compression process from the suction pressure to the economizer pressure; 502-502B is the mixing process of the vapor portion arrived via the suction inlet port and a vapor portion arrived via the economizer inlet port (thermodynamic state point 510); 502B-503 is the compression process from the economizer pressure to the discharge pressure; 503-504 is the cooling in the gas-cooler or the condensation in the condenser; 504-504′ is the additional cooling process in the traditional economized cycle; 504-505 is the additional cooling or sub-cooling process in the enhanced refrigerant system; 504′-504′A and 505-506 are the isenthalpic expansion processes in the traditional and enhanced cycles respectively; 504′A-501 and 506-507 are the evaporation processes in the traditional and enhanced cycles respectively; 508-509 is the pumping of liquid refrigerant process in the enhanced cycle; 509-510 is the evaporation process of the economizer branch refrigerant stream.

Thermodynamic state 509 associated with the refrigerant flowing through the economizer branch 107 of the enhanced refrigerant system is at the inlet to the economizer heat exchanger 103. Thermodynamic state 504B associated with the refrigerant flowing through the economizer branch 107 is at the inlet to the economizer heat exchanger of the traditional economized system. State 510 associated with the refrigerant flowing through the economizer branch 107 of the enhanced refrigerant system is at the outlet from the refrigerant conduit 103b of the economizer heat exchanger 103. The difference of refrigerant enthalpies in the thermodynamic state 510 and 509 is a heat transfer rate in the economizer heat exchanger between the economizer branch and the main refrigerant circuit of the enhanced system. The difference of refrigerant enthalpies in the thermodynamic state 510 and in state 504B is a heat transfer rate in the economizer heat exchanger between the economizer branch and the main refrigerant circuit in the traditional system. Obviously, that the enhanced cycle has higher cooling effect potential in the economizer heat exchanger since the refrigerant enthalpy in the thermodynamic state 509 is lower than the refrigerant enthalpy in the thermodynamic state 504B.

The diagram shows that the higher the economizer pressure is, the higher the heat transfer rate in the economizer heat exchanger is, for the enhanced system. On other hand, the higher this heat transfer rate is, the higher the cooling and heating capacities are. In addition, the higher the economizer pressure is, the lower the compressor power is. Therefore, the higher the economizer pressure is, the better the performance characteristics of the enhanced system, as has been already mentioned during discussion of FIGS. 3 and 4.

The higher the discharge pressure is, the higher the potential to increase the economizer pressure and improve performance of the enhanced system, with respect to the traditional system.

The diagram of FIG. 5 is exemplary and implies isobaric processes in all components, besides the compressor, the pump, and the expansion devices. However, due to pressure drop, all the discharge pressure states downstream of the compressor are gradually reduced, all the suction pressure states upstream to the compressor suction are gradually increased, and all the economizer pressure states upstream to the pump discharge are gradually increased.

The enhancement of the system performance using pumping of liquid refrigerant from the suction accumulator includes two unique features: 1) when the ambient temperature is reduced, in the heating mode of operation, the refrigerant density at the pump inlet is increased together with the pumping capacity; 2) when the ambient temperature is elevated, in the cooling mode of operation, the economizer pressure and the heat transfer rate in the economizer heat exchanger are increased as well. The first feature is completely opposite to the fact that vapor density is reduced at the compressor suction when ambient temperature decreases, degrading heating capacity and COP of the traditional economized systems. The second feature reduces a negative impact on performance characteristics of the enhanced refrigerant systems at elevated ambient temperatures.

FIG. 6 is boiling point elevation demonstration 600 that mitigates risks associated with potential cavitation phenomenon occurrence in the liquid refrigerant pump. As shown on FIG. 6, Xoil is the mass concentration of oil and Xref is the refrigerant quality at the inlet to expansion device, at the inlet to evaporator, and at the outlet of evaporator. When liquid refrigerant boils out, a mass concentration of oil in the remaining liquid portion of the refrigerant is increased, together with a boiling point of the oil-refrigerant mixture. The difference between the boiling point of an oil-refrigerant mixture and the evaporation temperature of a pure refrigerant is referred to as a boiling point elevation. The higher the oil concentration is, the higher the boiling point elevation is. The boiling point elevation acts as a sub-cooling and protects the pump from the cavitation that could potentially impact liquid pump reliability.

FIG. 6 illustrates a boiling point elevation demonstration 600; that if we have 2% of oil carryover in the compressor unit and refrigerant quality at the evaporator inlet is 0.2, then at vapor quality of 0.95 at the evaporator outlet, we will have 40% of oil in the oil-refrigerant mixture in the suction accumulator 106, and consequently at the inlet to the liquid pump 314. This provides sufficient boiling point elevation to avoid cavitation.

FIG. 7A illustrates an enhanced refrigerant system 700 including sequential multi-stage compression. The compressor unit 101 shown in FIG. 7A is a multi-stage compression device, which consists of a low-pressure compressor 719 and a high-pressure compressor 720. Each of these compressors has at least one compression stage. Each compression stage may have a plurality of parallel, or so-called tandem, compressors. The economizer inlet port 101a is located between these compressors 719 and 720. FIG. 7A shows the compressor unit 101 as a single device, however, the low-pressure compressor 719 and the high-pressure compressor 720 may represent separate compressor units.

FIG. 7B illustrates an enhanced refrigerant system 710 including parallel multi-stage compression. The compressor unit 101 shown in FIG. 7B is made of two parallel compressors: a high-pressure-ratio compressor 719 and a low-pressure-ratio compressor 720. The high-pressure-ratio compressor is associated with and operates between the suction inlet port and the discharge outlet port. The low-pressure-ratio compressor is associated with and operates between the economizer inlet port 101a and the discharge outlet port. Each of these compressors has at least one compression stage. Once again, each compression stage may have a plurality of parallel compressors. FIG. 7B shows the compressor unit 101 as a single unit, however, the high-pressure-ratio compressor 719 and the low-pressure-ratio compressor 720 may be constructed as separate compressor units. The compressors 719 and 720 in FIG. 7A and FIG. 7B may be equipped with oil separators returning oil back to enable better lubrication of moving parts of those compressors.

FIG. 8A illustrates an enhanced refrigerant system 800 charged and operating with a trans-critical refrigerant. The arrangement of FIG. 8A is applicable only for systems operating with trans-critical refrigerants. This arrangement implies incorporation of the economizer inlet port 101a with a discharge port 101b of the compressor unit 101. The liquid refrigerant stream pumped by the liquid refrigerant pump 314 from the suction accumulator 106 completely evaporates in the refrigerant conduit 103b of the economizer heat exchanger 103. Then it is mixed with hot refrigerant vapor discharged from the compressor unit 101. Thus, the liquid refrigerant stream from the suction accumulator 106 is pumped in parallel to the refrigerant of the compressor unit 101.

The economizer inlet port 101a may be physically incorporated with the discharge port 101b or the economizer branch 107 may be connected to a discharge line 101c downstream to the discharge port 101b. Also, the economizing inlet port 101a may be incorporated with the inlet side 102a of the heat rejection unit 102.

Application of a traditional sub-critical refrigerant within such a system requires a vapor temperature at the outlet from the refrigerant conduit 103b of the economizer heat exchanger 103 to be lower than a condensing temperature, in order to utilize the latent heat of the liquid refrigerant stream in the refrigerant conduit 103b of the economizer heat exchanger 103. This is not possible, and the traditional sub-critical refrigerants are not applicable for this system arrangement.

FIG. 8B illustrates an enhanced refrigerant system 810 including a split economizer port 812. FIG. 8C illustrates an enhanced refrigerant system 820 including a liquid receiver 821 located downstream of the heat rejection unit 102, with respect to the refrigerant flow. FIG. 8D illustrates an enhanced refrigerant system 830 including a liquid receiver 821 located between a condenser 102a and a sub-cooler 102b.

The arrangement in FIG. 8B may be applicable for sub-critical and trans-critical refrigerants. This arrangement implies placing of the economizer inlet port 101a between outlet from the heat rejection unit 102 and inlet to the economizer heat exchanger 103. The liquid refrigerant stream pumped by the liquid refrigerant pump 314 from the suction accumulator 106 is heated in the refrigerant conduit 103b of the economizer heat exchanger 103. Then it is mixed with refrigerant stream leaving the heat rejection unit 102. The economizer port 101a may be fabricated as a separate device. Also, it may be incorporated into the heat rejection unit 102 or into the economizer heat exchanger 103. However, this arrangement has a disadvantage in that it does not utilize any latent heat of the economizer stream.

If traditional sub-critical refrigerants are utilized, a liquid refrigerant receiver 821 may be installed at the outlet of the condenser 102 as shown on FIG. 8C. Also, the condenser may be split into two portions: the condensing portion 102a and sub-cooling portion 102b as shown on FIG. 8D. In that case the receiver 821 is installed between these portions. If the receiver 821 is applied to the arrangement shown in FIG. 7B, the economizing port 101a may be incorporated into the receiver.

In some embodiments, the compressor is a variable speed compressor. In some embodiments, the compressor is a multi-speed compressor. In some embodiments, the liquid refrigerant pump is a variable speed pump. In some embodiments, the refrigerant pump is a multi-speed pump.

FIG. 9 illustrates an enhanced heat pump 900. The heat pump consists of a refrigerant closed-loop circuit including the following components connected in sequence: the compressor unit 101 with the economizer inlet port 101a, the four-way reversing valve 209, the indoor heat exchanger unit 210 operating as a condenser or as a gas cooler, the expansion device 211, the economizer heat exchanger 103, the expansion device 212, the outdoor heat exchanger unit 213 operating as an evaporator, and the suction accumulator 106. The heat pump has the economizer branch 107 leading from the suction accumulator 106, through the refrigerant conduit 103b of the economizer heat exchanger 103, to the economizer inlet port 101a to the compressor unit 101.

In the heating mode, the four-way reversing valve 209 enables operation of the outdoor heat exchange unit 213 as an evaporator and the indoor heat exchange unit 210 as a condenser or as a gas cooler. The economizer heat exchanger 103 operates as a counterflow heat exchanger.

In the cooling mode, the four-way reversing valve 209 enables operation of the indoor heat exchange unit 210 as an evaporator and the outdoor heat exchange unit 213 as a condenser or a gas cooler. The economizer heat exchanger 103 operates as a parallel flow heat exchanger.

FIG. 10 illustrates the enhanced heat pump 1000 with a reversed flow arrangement in the economizer heat exchanger. If the counterflow arrangement is more effective for the cooling mode of operation, rather than for the heating mode of operation, then the economizer branch 107, and the refrigerant conduits 103a and 103b of the economizer heat exchanger 103 are connected as shown in FIG. 10. Also, it might be appropriate to have a crossflow arrangement in the economizer heat exchanger 103, to balance requirements in the heating and cooling modes of operation.

FIG. 11 illustrates the enhanced heat pump 1100 with a flash tank. The economizer heat exchanger 103 may be replaced with a flash tank 1116 of FIG. 11. The flash tank 1116 is a heat exchanger device providing a direct thermal contact between the refrigerant stream in the economizer branch 107 and the refrigerant stream in the main refrigerant circuit.

The flash tank 1116 consists of an inlet port 1116a, an outlet port 1116b, and two bi-flow ports 1116c and 1116d. The inlet port 1116a of the economizing branch 107 is fluidly associated with the liquid refrigerant pump 314. The outlet port 1116b of the economizer branch 107 is connected to the economizer inlet port 101a of the compressor 101. The bi-flow port 1116c and the indoor heat exchanger unit 210 are connected via an expansion device 1117; the bi-flow port 1116d is connected via expansion device 1118 to the outdoor heat exchanger unit 213 via an expansion device 1118.

In the heating mode, the flash tank 1116 is fed with a liquid refrigerant stream from the indoor heat exchange unit 210 via the expansion device 1117 and bi-flow liquid port 1116c. The outdoor heat exchange unit 213 is fed by from the flash tank 1116 via the bi-flow liquid port 1116d and the expansion device 1118.

In the cooling mode, the flash tank 1116 is fed with a liquid refrigerant stream from the outdoor heat exchanger unit 213 via the expansion device 1118 and bi-flow port 1116d. The indoor heat exchanger unit 210 is fed by from the flash tank 1116 via the bi-flow liquid port 1116c and the expansion device 1117.

If the system is designed to operate in one mode only, either cooling or heating, then the ports that are associated with the expansion devices 1117 and 1118 are not designed for the bi-flow operation and the four-way reversing valve is no longer needed.

When applicable, the design options for FIGS. 7A, 7B, 8A, 8B, and 8C are compatible to FIGS. 9-11. Also, the design options for FIGS. 9-11 are compatible to FIGS. 7A, 7B, 8A, 8B, 8C, and 8D.

Thus, the enhanced refrigerant system includes the following advantages in comparison with the traditional economized systems: 1) a portion of the total mass flow is pumped through the liquid refrigerant pump 314, which requires substantially lower power input; 2) the total mass flow is pumped through the evaporator unit, increasing the evaporator capacity; 3) the higher the economizing pressure is, the better cooling and heating capacities, compressor power, and COP are; 4) when ambient temperature decreases, density of liquid refrigerant at the pump inlet is increased, together with the pumping capacity, which works against degradation of heating capacity and COP; and 5) when ambient temperature increases, the economizer pressure increases, diminishing the extent of degradation of cooling capacity and COP.

As a result, the enhanced refrigerant system provides enhance heating and cooling capacities and heating and cooling COPs and provide the enhanced performance over a wider range of operating conditions than the traditional economized systems are designed for.

While some embodiments of the present invention have been disclosed in detail, it is to be understood that various modifications in its structure may be adopted without departing from the spirit of the invention or the scope of the following claims.

Claims

1. An enhanced refrigerant system comprising:

a refrigerant closed-loop circuit including a compressor unit, a heat rejection unit, an economizer heat exchanger, an expansion device, an evaporator unit, and a suction accumulator, said suction accumulator including an inlet, a vapor outlet, and a liquid outlet, said compressor unit including a suction inlet port, an economizer inlet port, and a discharge outlet port;
an economizer branch providing passage for a refrigerant stream flowing between said liquid outlet of said suction accumulator and said economizer inlet port of said compressor unit, said economizer branch including a liquid refrigerant pump and an economizer heat exchanger;
said economizer heat exchanger providing heat transfer interaction between said refrigerant stream in said economizer branch and said refrigerant stream in said refrigerant circuit;
said evaporator unit being configured and operated to provide refrigerant at its exit such that at least a portion of this refrigerant is a liquid and non-evaporated phase; and
wherein said liquid refrigerant pump of said economizer branch pumps said liquid refrigerant from the exit of said evaporator unit through said economizer branch and said economizer heat exchanger, and where at least a portion of said liquid refrigerant evaporates and forms a vapor refrigerant stream in said economizer branch that flows into said economizer inlet port of said compressor unit.

2. An enhanced refrigerant system of claim 1 wherein said enhanced refrigerant system operates in a sub-critical cycle and said heat rejection unit functions as a condenser.

3. An enhanced refrigerant system of claim 1 wherein said enhanced refrigerant system operates in a trans-critical cycle and said heat rejection unit functions as a gas cooler.

4. An enhanced refrigerant system of claim 1 that includes at least one of:

(a) said heat rejection unit including a plurality of individual heat rejection units.
(b) said expansion device including a plurality of individual expansion devices.
(c) said evaporator unit including a plurality of individual evaporators.
(d) said suction accumulator including a plurality of individual suction accumulators.
(e) said liquid refrigerant pump includes a plurality of individual pumps; and
(f) said economizer heat exchanger including a plurality of individual heat exchangers.

5. An enhanced refrigerant system as recited in claim 1 wherein said economizer heat exchanger includes a plurality of channels.

6. An enhanced refrigerant system as recited in claim 1 wherein said heat rejection unit is an outdoor unit and said evaporator unit is an indoor unit.

7. An enhanced refrigerant system of claim 1 wherein said heat rejection unit is an indoor unit and said evaporator unit is an outdoor unit.

8. An enhanced refrigerant system of claim 6 wherein said refrigerant closed-loop circuit includes a four-way reversing valve, said four-way reversing valve includes a vapor inlet port, a vapor outlet port, a first bi-flow port, and a second bi-flow port, said vapor inlet port is connected to said discharge port of said compressor unit, said vapor outlet port is connected to said vapor inlet port and to said suction accumulator, said first bi-flow port is connected to said refrigerant closed-loop circuit at a location upstream of said outdoor unit, and said second bi-flow port is connected to said refrigerant closed-loop circuit at a location downstream of said indoor unit.

9. An enhanced refrigerant system of claim 1 wherein said compressor unit and said heat rejection unit are combined into a single unit assembly.

10. An enhanced refrigerant system of claim 1 wherein said expansion device is combined with said evaporator unit.

11. An enhanced refrigerant system of claim 1 wherein said liquid refrigerant pump and said suction accumulator are combined into a single unit assembly.

12. An enhanced refrigerant system as recited in claim 1 wherein said compressor unit includes a low-pressure compressor, a high-pressure compressor, and said economizer inlet port that is located between said low-pressure and said high-pressure compressor; each of said compressors including at least one compression stage.

13. An enhanced refrigerant system as recited in claim 12 wherein at least one of said at least one compression stage includes a plurality of parallel compressors.

14. An enhanced refrigerant system of claim 12 wherein said low-pressure compressor and said high-pressure compressor are separate compressor units.

15. An enhanced refrigerant system as recited in claim 1 wherein said compressor unit includes two parallel compressors, a first compressor provides a refrigerant passage between said economizer inlet port and said discharge outlet port, and a second compressor provides a refrigerant passage between said suction inlet port and said discharge outlet port; each of said compressors includes at least one compression stage.

16. An enhanced refrigerant system as recited in claim 15 wherein said first compressor includes a plurality of parallel compressors.

17. An enhanced refrigerant system as recited in claim 15 wherein said second compressor includes a plurality of parallel compressors.

18. An enhanced refrigerant system of claim 15 wherein said first compressor and said second compressor are separate compressor units.

19. An enhanced refrigerant system of claim 1 wherein said economizer heat exchanger is a counterflow heat exchanger in a primary operational mode.

20. An enhanced refrigerant system of claim 1 wherein said economizer heat exchanger is a parallel flow heat exchanger in a primary operational mode.

21. An enhanced refrigerant system of claim 1 wherein said economizer heat exchanger is a crossflow heat exchanger.

22. An enhanced refrigerant system of claim 1 wherein said economizer heat exchanger is a flash tank having a vapor inlet port, a vapor outlet port, and two liquid ports and providing a direct thermal contact between refrigerant stream in said refrigerant closed-loop circuit and refrigerant stream in said economizer branch.

23. An enhanced refrigerant system of claim 22 wherein at least one of two said liquid ports has an expansion device for said refrigerant closed-loop circuit upstream of said flash tank.

24. An enhanced refrigerant system of claim 23 wherein at least one of two said liquid ports and said flash tank are combined into a single unit.

25. An enhanced refrigerant system of claim 1 wherein said enhanced refrigerant system operates in a trans-critical cycle and said economizer inlet port of said compressor unit is combined with said discharge port of said compressor unit.

26. An enhanced refrigerant system of claim 1 wherein said enhanced refrigerant system operates in a trans-critical state and said economizer inlet port of said compressor unit is a separate device located downstream of said heat rejection unit and upstream of said economizer heat exchanger.

27. An enhanced refrigerant system of claim 1 wherein said compressor is a variable speed compressor.

28. An enhanced refrigerant system of claim 1 wherein said compressor is a multi-speed compressor.

29. An enhanced refrigerant system of claim 1 wherein said liquid refrigerant pump is a variable speed pump.

30. An enhanced refrigerant system of claim 1 wherein said liquid refrigerant pump is a multi-speed pump.

31. An enhanced refrigerant system of claim 2 wherein a liquid refrigerant receiver is installed at a location downstream said heat rejection unit.

32. An enhanced refrigerant system comprising:

an economizer branch providing passage of a refrigerant stream flowing between a liquid outlet of a suction accumulator and an economizer inlet port of a compressor unit, said economizer branch including a liquid refrigerant pump and configured to pass through an economizer heat exchanger that provides for heat transfer interaction between said refrigerant stream in said economizer branch and a refrigerant stream in a main refrigerant closed-loop circuit; and
wherein said main refrigerant closed-loop circuit also passes through said economizer heat exchanger and includes a compressor unit including a suction inlet port that is located downstream of said suction accumulator.

33. The enhanced refrigerant system of claim 32 wherein said main refrigerant closed-loop circuit further includes an evaporator unit being configured and operated to provide refrigerant at its exit such that at least a portion of this refrigerant is in a liquid and non-evaporated phase and wherein said liquid refrigerant pump of said economizer branch pumps said liquid refrigerant through said economizer branch and into said economizer heat exchanger in heat transfer interaction with the refrigerant flowing through said main closed-loop refrigerant circuit, and where at least a portion of said liquid refrigerant evaporates and forms a vapor refrigerant stream in said economizer branch that flows into said economizer inlet port of said compressor unit.

34. An enhanced refrigerant system as recited in claim 33 wherein said compressor is a variable speed compressor.

35. An enhanced refrigerant system as recited in claim 33 wherein said compressor is a multi-speed compressor.

36. An enhanced refrigerant system as recited in claim 33 wherein said liquid refrigerant pump is a variable speed pump.

37. An enhanced refrigerant system as recited in claim 33 wherein said liquid refrigerant pump is a multi-speed pump.

38. An enhanced refrigerant system as recited in claim 11 wherein said liquid refrigerant pump is located below said suction accumulator.

39. An enhanced refrigerant system as recited in claim 11 wherein said liquid refrigerant pump is located above said suction accumulator.

40. An enhanced refrigerant system as recited in claim 11 wherein said liquid refrigerant pump is located inside said suction accumulator.

41. An enhanced refrigerant system as recited in claim 11 wherein said liquid refrigerant pump is located below, above or inside of said accumulator.

42. An enhanced refrigerant system of claim 8 wherein said compressor unit, said four-way reversing valve, and said heat rejection unit are combined into a single unit.

43. An enhanced refrigerant system of claim 1 wherein said compressor unit is equipped with an oil separator returning oil to said compressor.

44. An enhanced refrigerant system as recited in claim 1 wherein said heat rejection unit is a mini-channel heat exchanger.

45. An enhanced refrigerant system as recited in claim 1 wherein said heat absorption unit is a mini-channel heat exchanger.

Patent History
Publication number: 20100005831
Type: Application
Filed: Feb 2, 2007
Publication Date: Jan 14, 2010
Applicant: CARRIER CORPORATION (Farmington, CT)
Inventors: Igor B. Vaisman (West Hartford, CT), Michael F. Taras (Fayetteville, NY), Joseph J. Sangiovanni (West Suffield, CT)
Application Number: 12/524,496
Classifications
Current U.S. Class: Lubricant Separator (62/470); Plural Compressors Or Multiple Effect Compression (62/510); Heat Exchange Between Diverse Function Elements (62/513)
International Classification: F25B 43/02 (20060101); F25B 1/10 (20060101); F25B 41/00 (20060101);