CONDENSATE-FREE OUTDOOR AIR COOLING UNIT

Abstract

A highly efficient condensate-free cooling unit functionally based on vapor-compression refrigeration cycle has been described. The condensate collected from the evaporator of the cooling unit is routed through a sub-cooling heat exchanger where it exchanges heat with the primary heat exchange medium emerging through the condenser of the cooling unit, thus, sub-cooling the primary heat exchange medium to a lower temperature before it enters the expansion valve. Emerging from the sub-cooling heat exchanger, the condensate flows through a condensate outlet pipe into multiple spray nozzles disposed over the condensate outlet pipe. The spray nozzles sprinkle the condensate over the hot air blown into the condenser to reduce its temperature. The cooling unit has a substantially higher coefficient of performance compared to the conventional cooling units utilizing vapor-compression refrigeration cycle, and eliminates the problems of condensate removal persistent in the art.

Description

FIELD OF THE INVENTION

The present disclosure deals generally with heat exchange devices, and more specifically with cooling devices utilizing the vapor compression refrigeration cycle.

BACKGROUND OF THE INVENTION

Vapor compression refrigeration cycles are widely used in many cooling systems for achieving refrigerating effects, including refrigerators, air-conditioning systems, industrial and commercial refrigeration systems and the like. Typical applications of such cooling units include warehouses, offices, private and public residential buildings, hospitals, hotels, restaurants and cafeteria. Succinctly, the process of refrigeration refers to extracting heat from a space and rejecting it elsewhere, thus lowering the temperature of the space. Vapor compression refrigeration cycles use a refrigerant for this process. The cycle includes four elementary components: a compressor, a condenser, an expansion valve, and an evaporator. The circulating refrigerant enters the compressor in the form of saturated vapor and undergoes isentropic compression, thus increasing its pressure and temperature, and converting into a superheated vapor. The refrigerant then enters the condenser, where it comes into thermal contact with cold water or air and rejects heat to it. In the process, the refrigerant absorbs heat (sensible heat) and is converted from superheated vapor to saturated vapor. After further absorbing latent heat, the saturated vapor eventually is converted to saturated liquid.

At the condenser exit, the refrigerant is thermodynamically a saturated liquid. That liquid passes through an expansion valve, where it expands, undergoing a reduction in pressure and consequently a partial flash evaporation. This process converts the refrigerant into a mixed liquid and vapor and reduces its temperature to a level below that of the space to be refrigerated. The mixture then enters the evaporator where it extracts latent heat from the space and thus completely vaporizes to a saturated vapor. The saturated vapor re-enters the compressor to complete the refrigeration cycle. This description pertains to an ideal vapor-compression refrigeration cycle, neglecting practical real-world effects such as the frictional pressure drop in the system and the slight thermodynamic irreversibility.

In the evaporator, the air to be conditioned passes over heat exchange tubes, exchanging heat with the refrigerant and simultaneously lowering the air temperature and vaporizing the refrigerant. The ambient air blown over the evaporator's heat exchange tubes generally carries a certain amount of water vapor, and in the course of exchanging heat with the refrigerant, this vapor partially condenses, forming a water condensate which starts dripping off from the evaporator.

Appropriate disposal of cooling unit condensate poses a challenging problem. Providing a separate piping system or condensate collection containers can be burdensome and expensive. Dehumidifiers are employed in certain applications, to reduce the moisture content of the air entering the evaporator. Sometimes, excessive extraction of heat from the air even causes frost or ice to build-up in the evaporator and the surface of its heat exchange tubes, causing the dehumidifier to stop. Thermostats have been used in the art to detect the frost accumulation conditions, by identifying moments when the air temperature drops below a certain value. The compressor then shuts down until the system warms up, but those actions shut down the cooling system itself. Efforts have been also made in the art for routing the condensate out of the evaporator casing through pumps.

Another challenge is increasing the operating efficiency of the cooling units. A measure of the operating efficiency of such cycles is the coefficient of performance (COP), which is the ratio of the obtainable useful refrigeration effect to the power required to drive the cycle (including the compressor driving power). Either an increase in the useful refrigeration effect or a decrease in the power input to the compressor may increase the operating efficiency. Substantially successful attempts in that direction include disposing an intercooler between the condenser and the expansion valve, routing the high pressure/high temperature refrigerant through the intercooler, where it exchanges heat with the low temperature refrigerant entering the evaporator. This decreases the enthalpy of the high pressure refrigerant and thus increases the coefficient of performance. To increase efficiency, the added compressor power input must be less than the increase in the useful refrigeration effect. In some applications, the high pressure value of the refrigerant in the cycle is varied and a corresponding change in the value of the COP is observed to derive a correlation. This correlation is then used to identify the high pressure value at which the COP maximizes. Many times, the analysis and identification of this correlation may be time consuming and inaccurate.

At present, there exists a need for an efficient cooling unit that could have a considerably higher coefficient of performance compared to the conventionally used systems, and would simultaneously address the problem of condensate removal from cooling units, in an effective manner.

SUMMARY

The present invention is directed to a highly effective and advantageous cooling unit functionally based on vapor compression refrigeration cycle. The cooling unit has an operable efficiency (coefficient of performance) substantially greater than the coefficient of performance of the conventionally used cooling units, including refrigerators and air-conditioning devices. It further effectively eliminates problem of condensate removal persistent in the conventional systems.

In one aspect, a cooling unit for cooling an ambient medium is provided that includes a compressor, a condenser, an expansion valve, an evaporator, a sub-cooling heat exchanger and a condensate pump. The cooling unit functionally utilizes vapor compression refrigeration cycle for achieving refrigerating effect. The evaporator includes a number of heat exchange tubes through which a primary heat exchange medium flows. The heat exchange tubes remain in continuous thermal communication with the ambient medium. The condensate pump is connected to the evaporator to route the condensate collected within the evaporator through the sub-cooling heat exchanger. The sub-cooling heat exchanger has a condensate inlet meant to receive the routed condensate and a condensate outlet for delivering the condensate to a number of spray nozzles connected to it through a condensate outlet pipe. The condensate outlet pipe has a set of perforations provided on its surface through which the sub-cooling heat exchanger remains in continuous fluid communication with the spray nozzles. Effectively, the condensate is routed from the evaporator to the sub-cooling heat exchanger and further to the spray nozzles. The spray nozzles are directly coupled with hot air blowing into the condenser and are meant to sprinkle the condensate thereon to reduce its temperature. The sub-cooling heat exchanger further has a primary heat exchange medium inlet and a primary heat exchange medium outlet for a continuous influx and efflux of the primary heat exchange medium through it. During operations, the condensate and the primary heat exchange medium remain in continuous thermal communication within the sub-cooling heat exchanger and exchange heat with each other.

The cooling unit has a considerably high coefficient of performance compared to the conventionally used cooling systems. Further, the cooling unit is effectively dry and perfectly addresses the problems of condensate removal persistent in the art.

Additional features and advantages of the invention will be made apparent from the following detailed description of illustrative embodiments that proceed with reference to the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

The summary above, as well as the following detailed description of preferred embodiments, is better understood when read in conjunction with the appended drawings. For the purpose of illustrating the invention, exemplary constructions of the invention are shown in the drawings. The invention is not limited to the specific methods and instrumentalities disclosed however. Moreover, those in the art will understand that the drawings are not to scale. Where possible, like elements are indicated by identical numbers.

FIG. 1 is the pressure-enthalpy chart for a refrigerant undergoing a conventional vapor compression refrigeration cycle.

FIG. 2 is a schematic diagram showing the basic components of a conventional vapor compression refrigeration system.

FIG. 3 illustrates the different components of an exemplary cooling unit in accordance with the present disclosure

FIG. 4 is another schematic diagram representing the different components of the cooling unit of FIG. 3

FIG. 5 shows the evaporator of the cooling unit of FIG. 4 in greater details.

FIG. 6 shows the pressure-enthalpy chart for the primary heat exchange medium circulating within the refrigeration cycle utilized by the exemplary cooling unit of FIG. 4 in accordance with the present disclosure.

FIG. 7 illustrates the heat exchange process within the condenser of the cooling unit of FIG. 4.

DETAILED DESCRIPTION OF ILLUSTRATIVE EMBODIMENTS

The description below illustrates embodiments of the claimed invention. This description discloses aspects of the invention but does not define or limit the invention, such definition and limitation being contained solely in the claims appended hereto. Those of skill in the art will understand that the invention can be implemented in a number of ways different from those set out here, in conjunction with other present or future technologies.

As used herein, the following terms carry the had indicated meanings: “Ambient medium” is the medium targeted for conditioning by the disclosed cooling unit. In a building air-conditioning system, for example, the ambient medium would be the air inside the building. “Primary heat exchange medium” designates the refrigerant. In most circumstances, the primary heat exchange medium would be a refrigerant capable of exchanging heat with the ambient medium targeted for conditioning.

FIG. 1 shows the pressure-enthalpy diagram for a refrigerant undergoing a typical single-stage vapor compression refrigeration cycle. The description should be read in conjunction with FIG. 2, which depicts the components of a conventional vapor-compression refrigeration system and the states corresponding to those marked in the pressure-enthalpy diagram for the refrigerant, as shown in FIG. 1. The refrigerant, in the form of saturated vapor at point 1 (pressure P₁ and Temperature T₁), enters the compressor where it is compressed to a high pressure P₂ and high temperature T₂, which converts the refrigerant from the thermodynamic state of saturated vapor to superheated vapor at point 2. At this point, the refrigerant exits the compressor. The increase in refrigerant pressure causes an increase in temperature, thus bringing it to the state 2 where it can be condensed by typically available air or water.

At state point 2, the refrigerant enters the condenser and flows through the condenser coils, where it comes in thermal contact with the air or water flowing across the coils. The coils allow the refrigerant to reject sensible heat to the air or water, converting it from superheated vapor to saturated vapor in the course of traversing from point 2 to point 2. The rejected heat is called sensible heat because its loss leads to a change in refrigerant temperature during the conversion from superheated to saturated vapor. The enthalpy of the refrigerant decreases in this process as heat is extracted, or equivalently, heat is rejected by it to the air or water in contact. The refrigerant absorbs further heat, and loses further enthalpy, until it is converted to saturated liquid at point 3.

At point 3, the refrigerant exits the condenser and enters the expansion valve. Here, saturated liquid refrigerant expands, with a sharp loss of pressure, to point 4 corresponding to the exit from the expansion valve. This sudden decrease in pressure from a value P₃ to P₄ causes a corresponding drop in the temperature of the liquid refrigerant from the corresponding value T₃ to T₄ and a portion of the liquid refrigerant flashes off into vapor, leaving the refrigerant in mixed liquid and vapor form at the exit 4 of the expansion valve. The enthalpy of the refrigerant however remains constant from point 3 to 4, as shown in FIG. 1, as no external heat is added or subtracted.

At point 4, the refrigerant enters the evaporator where it comes into contact with the air to be cooled. This air flows across the evaporator coils and exchanges heat with the refrigerant flowing in the coils. By extracting the latent heat of vaporization from the air, the refrigerant is converted from a mixed state to saturated vapor at point 1, where it again enters the compressor and completes the refrigeration cycle.

The coefficient of performance (COP) of a vapor compression refrigeration cycle measures the effectiveness of the cycle and is defined as the ratio of the useful refrigerating effect to the total power required to operate the cycle. For the case of a mechanical vapor compression refrigeration cycle, the total power required is usually consumed in driving the mechanical components including the compressor, the fans, the pumps etc. Mathematically:

$\begin{array}{cc}\mathrm{COP}=\frac{Q_{\mathrm{evaporator}}}{W_{\mathrm{net}}}& \left(1\right)\end{array}$

• • where, Q_evaporator=net rate of heat extracted by the evaporator from the ambient medium to be conditioned; and
  • W_net=Net power required to drive the cycle components, including the compressor, the fans and the pumps

As seen from equation (1) above, there are two possible ways to increase the COP of a vapor compression refrigeration cycle, one by increasing the rate of energy extraction from the ambient medium to be conditioned (Q_evap) and the other by decreasing the power input to the mechanical components, mainly the compressor (W_net).

As shown in FIG. 1, if points 1 and 2 represent the states of the refrigerant during its entry into and exit from the compressor, respectively, and h₁ and h₂ represent the specific enthalpies (enthalpy per unit mass) of the refrigerant at states 1 and 2 respectively, and assuming the process 1→2 being reversible and adiabatic, then the amount of power required to drive the compressor is:

W _net =m(h₂−h₁)   (2)

• • where m=mass flow rate of the refrigerant at the inlet to the compressor

The quantity Q_evap in eq. (1) depends on the temperature of the ambient air flowing into the evaporator, which is uncontrollable to a certain extent, at a specific mass flow rate. Further, the magnitude of the higher refrigerant pressure value (P₂) corresponding to point 2 in FIG. 1, depends upon the temperature of the air entering the condenser. The purpose of compressing the refrigerant is to increase its pressure, and hence its temperature, to a value where it can easily reject heat to the medium entering the condenser. In that sense, the refrigerant has to be compressed to an extent where the temperature of superheated vapor refrigerant (T₂) is greater than the temperature of the condenser inlet air. If somehow, this extent of compression is reduced (to a point somewhere between states 1 and 2), to the point A or point B as represented in FIG. 1 for instance, the power required to drive the compressor would decrease consequently, leading to a tremendous increase in the coefficient of performance of the refrigeration cycle. Specifically, in a case where the refrigerant is compressed to a point A, instead of point 2, the approximate power input to the compressor would be:

W ^/ _net =m(h_a−h₁)   (3)

• • where h_a=specific enthalpy of refrigerant at A and h_a<h₂,W^/_net<W _net

The present disclosure uses this concept to substantially increase the coefficient of performance of a cooling unit while also effectively eliminating the problem of condensate removal.

FIG. 3 is a schematic diagram representing the components of cooling unit 300 in accordance with the present disclosure, as well as the heat exchange process at the evaporator and the condenser side. FIG. 4 further shows the different labeled components of the cooling unit shown in FIG. 3. Explaining in conjunction with these figures, cooling unit 300 includes an evaporator 410, a compressor 420, a condenser 430, a sub-cooling heat exchanger 450, an expansion valve 440, a condensate pump 460 and a number of spray nozzles 480. The primary heat exchange medium flows along the circuit represented by solid thick lines shown in FIG. 3 and FIG. 4, while condensate flows along the dotted-line circuit.

As illustrated, the primary heat exchange medium (refrigerant), in the form of saturated vapor, enters the compressor 420 at state 1 and converts into superheated vapor at point 2^/. The superheated vapor refrigerant rejects heat to air 490 blown into the condenser 430, and as its temperature drops it converts first into saturated vapor and finally into saturated liquid at the exit 3^/ from the condenser 430. The refrigerant is then routed through a sub-cooling heat exchanger 450, where it exchanges heat with condensate 492 collected from the evaporator 410 through the condensate pump 460. Thus, the refrigerant temperature further decreases and it converts from saturated liquid to sub-cooled liquid at point 3^//. The condensate handling system is discussed in detail below. Exiting from the sub-cooling heat exchanger, the refrigerant enters expansion valve 440, where it expands and undergoes an abrupt decrease in pressure, thus lowering its temperature. A fraction of the liquid refrigerant undergoes partial flash evaporation, and it exits expansion valve 440 as a mixed liquid and vapor at point 4^/. Entering the evaporator 410, the refrigerant, its temperature substantially lowered, extracts heat from ambient air 495, and vaporizes completely before exiting the evaporator 410 at state 1. The vaporized refrigerant reenters the compressor 420 to complete the refrigeration cycle.

The ambient air 495 generally has a certain absolute humidity, and the heat extraction process generally results in at least a portion of that moisture condensing as condensate 492 within the evaporator 410. Condensate pump 460 routes this condensate 492 along a condensate flow path K→L→M as represented by the dotted lines in FIG. 4. Following this path, the condensate 492 enters the sub-cooling heat exchanger 450 through a condensate inlet 452, exchanges heat with the refrigerant flowing through the sub-cooling exchanger 450, and finally exits through a condensate outlet 454. Thereafter, the condensate 492 flows out through a condensate outlet pipe 485 connected to the condensate outlet 454. A number of spray nozzles 480 are mounted spatially equidistant on the condensate outlet pipe 485, and set of perforations 486 convey condensate 492 to the spray nozzles 480.

The spray nozzles 480 are arranged to continuously spray condensate 492 into the air stream 490 flowing into condenser 430. That flow can be assisted by a blower fan 470, and the condensate spray acts on air stream 490 to reduce its temperature, as is clear to those of skill in the art. As a result, air stream 490 makes contact with the condenser coils at a reduced temperature, enhancing its ability to extract heat from the refrigerant.

Compressor 420 could be any suitable compressor known in the art usable for refrigerating units, and the selection of the exact compressor type is based on the operating conditions of the cooling unit 300 and certain associated design and maintenance criteria. To minimize losses in the refrigerant pressure and reduce the maintenance activities, and for cases where prolonged maintenance-free operations of the cooling unit 300 is desirable, a hermetically sealed or a semi-hermetic compressor would be preferred. An electric motor (not shown) is used a source of continuous power supply to the compressor 420 during operating conditions. Further, compressor 420 can be either a positive displacement or dynamic compressor depending upon the cooling load required and the environment of installation of the cooling unit 300. In large buildings and warehouses demanding huge cooling capacity, a dynamic type single-stage centrifugal compression would be preferred. Cases wherein extremely high output pressures at the compressor outlet are required, a multi-staged centrifugal compressor would be preferably used. In certain embodiments, a variable displacement type compressor may be used wherein the mass flow rate of the refrigerant needs to be frequently controlled based on variations in the temperature of the air blown into the evaporator 410.

Any suitable refrigerant known to those in the art can be used as the primary heat exchange medium for circulation within the cooling unit 300 for practicing the present disclosure. Common examples include the well-known family of refrigerants denoted by the ‘R-number’ system including R-11, R-22, R-134 (a) and others.

The sub-cooling heat exchanger 450 is be designed to provide sufficient capacity to incorporate high volume of condensate 492 generated during prolonged operations of the cooling unit 300. Those skilled in the art would recognize that the size and capacity of the sub-cooling heat exchanger 450 can be varied based on the desired cooling capacity of the cooling unit 300, as well as upon the expected conditions of ambient air. During inoperative conditions of the cooling unit 300, the sub-cooling heat exchanger 450 would act as a reservoir for the condensate 492 collected during former operations. Further, the aperture of the condensate outlet 454 is preferably smaller than the aperture of the condensate inlet 452, so that the volume of condensate entering the sub-cooling heat exchanger 450 is more than the volume leaving it. In that manner, a certain volume of condensate accumulates in the sub-cooling heat exchanger 450, and thus the level of condensate rises during continuing operations of the cooling unit 300. The condensate inlet 452 is provided at a higher elevation with respect to the condensate outlet 454 to ensure that there is a continuous influx of the condensate 492 within the sub-cooling heat exchanger 450. Eventually, with the elevated level of the accumulated condensate 492 in the sub-cooling heat exchanger 450, the condensate 492 flows out through the condensate outlet 454 into the condensate outlet pipe 485 by virtue of the achieved velocity of efflux. Further, the pressurized influx of the condensate 492 into the sub-cooling heat exchanger 450 by the condensate pump 460 provides additional expelling impulse for the volume of condensate 492 leaving the sub-cooling heat exchanger through the condensate outlet 454.

Condensate pump 460 pressurizes the condensate 492 to flow through the sub-cooling heat exchanger 450 and on to spray nozzles 480. The condensate velocity in that flow circuit should be sufficient to allow the spray nozzles 480 to discharge the condensate as a spray of fine droplets, maximizing heat exchange between the condensate and the air stream 490. To minimize energy consumption, hydraulic spray nozzles are preferred. However, those skilled in the art would understand that any other type of spray nozzles conventionally known in the art may also be used, including gas atomized spray nozzles, thus, not limiting the scope of the invention.

The sub-cooling heat exchanger 450 further includes a primary heat exchange medium inlet 455 and a primary heat exchange medium outlet 456 for a continuous influx and discharge of refrigerant. Exiting the condenser 430, the refrigerant enters the sub-cooling heat exchanger 450 inlet 455, flows through a set of cooling coils 458, and finally emerges through outlet 456. That flow allows the refrigerant to reject heat to the condensate 492, and to emerge at point 3^// as a sub-cooled liquid. This decreases the temperature of the refrigerant as it exits the sub-cooling heat exchanger 450. This heat exchange is an isobaric process, occurring at the constant saturation pressure of the saturated liquid refrigerant exiting the condenser 430 at point 3^/. The specific enthalpy of the refrigerant decreases as it moves from the state 3^/ to the state 3^//.

FIG. 5 illustrates the evaporator 410 to be used in the cooling unit 300 of FIG. 4, in further detail. Evaporator 410 includes a casing 501 having an anterior surface 502 and a posterior surface 503. A set of heat exchange tubes 504 are disposed in parallel within the casing 501. The heat exchange tubes 504 extend along the entire length of the casing 501, lying substantially perpendicular to the anterior and posterior surfaces 502 and 503 respectively. Further, the anterior surface 502 and posterior surface 503 have perforations through which the heat exchange tubes extend into and out of casing 501. Fins 505 are equidistantly disposed within the casing over the heat exchange tubes 504. Those skilled in the art would recognize that different evaporator structures could easily be substituted for the disclosed component without altering the performance of the disclosed embodiment.

FIG. 6 illustrates the pressure-enthalpy diagram for the refrigerant circulating in the cooling unit 300 of the present disclosure. The solid lined path 1→2^/→2^//→3^/→3^//→4^/>1 represents the respective states marked in FIG. 3 and FIG. 4, for the refrigerant flowing in the cooling unit 300 according to the present disclosure. Path 1→2^/→2→3→4→1 (dotted lines) represents the pressure-enthalpy variation of the refrigerant if it flows through a conventional cooling unit based on the vapor-compression refrigeration cycle and not equipped with a sub-cooling heat exchanger and spray nozzles, as shown by FIG. 2. The horizontal axis OX represents the specific enthalpy (enthalpy per unit mass) of the refrigerant at different thermodynamic state points, while the vertical axis OB represents the pressure of the refrigerant at different states. The two superimposed paths lay out a comparison between conventional cooling units and the inventive cooling unit 300 of FIG. 4.

As shown, in conventional refrigerating units, the air at the inlet to the condenser is at a relatively higher temperature, and hence the refrigerant is compressed to a higher pressure P₂ and correspondingly a higher temperature T₂, so that it can easily reject heat to the air entering the condenser. In the inventive cooling unit 300, the air blown into the condenser 430 is sprinkled with condensate 492 by the spray nozzles 480 (FIG. 4), before that air stream enters the condenser 430. Hence, the temperature of the air 490 entering the condenser 430 is relatively lower. Therefore, the refrigerant in the cooling unit 300 needs to be compressed to a lower pressure represented by a point 2^/ lying somewhere on the curve joining the points 1 and 2 in FIG. 6. The exact location of this point 2^/ on the curved line 1→2 depends on certain factors, the primary ones being the temperature of the air 490 entering the condenser 430, its mass flow rate and the mass flow rate of the refrigerant circulating in the cooling unit 300.

Further, during process 3^/→3^//, the refrigerant flows through the sub-cooling heat exchanger where it rejects heat to the condensate and emerges as a sub-cooled liquid at point 3^//. Process 3^//→4^/ represents the expansion of the refrigerant within the expansion valve 440. This expansion abruptly decreases the pressure of the refrigerant from P₂^/ to P₁, though the enthalpy of the refrigerant remains at a constant value h₄^/ during the process. In comparison, the refrigerant in conventional cooling unit of FIG. 2 would undergo the path 2→3 in the condenser, emerge as a saturated liquid at point 3, and finally expand in the expansion valve isobarically to the point 4 (process 3→4 in FIG. 6), consequently decreasing its pressure from P_{2 to P1.}

Portion 510 on axis OX in FIG. 6 represents the decrease in the required power input to the compressor achieved by the inventive cooling unit 300 of FIG. 4, in comparison to a conventional cooling unit of FIG. 2. As set out earlier in Eq. (2), the power required to drive the compressor of a conventional cooling unit of FIG. 2 is:

W _net ={dot over (m)}(h₂−h₁)   (2)

• • where, {dot over (m)}=mass flow rate of the refrigerant flowing in the vapor-compression refrigeration cycle.

Using the same refrigerant, at the same mass flow rate {dot over (m)}, the power required to drive the compressor of the inventive cooling unit of FIG. 4 is:

W ^/ _net ={dot over (m)}(h₂^/−h₁)   (3)

Since h₂^/<h₂ (as seen in FIG. 5), W^/_net<W_net

The decreased power input to the compressor, as achieved by the cooling unit 300 of FIG. 4 is obtainable by subtracting Eq. (3) from Eq. (2), and is:

ΔW=W_net−W^/_net={dot over (m)}(h₂−h₂^/)   (4)

The length of the portion 510 on the enthalpy axis OX of FIG. 5 actually represents this decreased compressor power input per unit mass flow rate of the refrigerant (ΔW/{dot over (m)}), as achieved by the cooling unit 300 of FIG. 4.

Further, portion 520 in FIG. 6 represents the enhanced cooling effect achieved by the cooling unit 300 of FIG. 4. This enhanced cooling effect 520 is equivalent to the additional heat energy extracted by the evaporator 410 of cooling unit 300 (denoted by Q^/_evap) from the ambient air 495 to be conditioned, in comparison to the heat actually extracted by the evaporator of a conventional cooling unit of FIG. 2 (denoted by Q_evap).

For a conventional cooling unit of FIG. 2, the refrigerant undergoes the process 4→1 (shown in FIG. 6) in the evaporator and hence, the heat gained by the refrigerant from the ambient air to be conditioned is:

Q _evap ={dot over (m)}(h₁−h₄)   (5)

In the inventive cooling unit 300, the refrigerant is further sub-cooled to state 3^// and hence undergoes the process represented by 4^/→1 within the evaporator. Therefore, the heat gained by the refrigerant is:

Q ^/ _evap ={dot over (m)}(h₁−h₄^/)   (6)

Since, h₄^/<h₄, we have Q^/_evap>Q_evap

The approximate enhanced cooling effect gained by the cooling unit 300, can be obtained from Eq. (5) and Eq. (6) above, and is:

ΔQ_evap=Q^/_evap−Q_evap={dot over (m)}(h₄−h₄^/)   (7)

The length of the portion 520 on the enthalpy axis OX of FIG. 5 actually represents this enhanced cooling effect per unit mass flow rate of the refrigerant, i.e., ΔQ_evap/{dot over (m)}.

The coefficient of performance of the cooling unit 300 of FIG. 4 is substantially increased in this manner, compared with the COP of the conventional cooling units utilizing vapor-compression refrigeration cycles. As stated in eq. (i) earlier, the COP of a refrigeration cycle is given by:

$\begin{array}{cc}\mathrm{COP}=\frac{Q_{\mathrm{evaporator}}}{W_{\mathrm{net}}}& \left(1\right)\end{array}$

Referring to FIG. 6, the value of COP for a conventional unit of FIG. 3 would be:

$\begin{array}{cc}\begin{array}{c}\mathrm{COP}=\frac{Q_{\mathrm{evaporator}}}{W_{\mathrm{net}}}\\ =\frac{\left(h_1-h_4\right)}{\left(h_2-h_1\right)}\end{array}& \left(8\right)\left[\mathrm{from}\mathrm{eq}\left(2\right)\mathrm{and}\left(5\right)\right]\end{array}$

For the inventive cooling unit 300 of FIG. 4, the enhanced COP is given by:

$\begin{array}{cc}{\mathrm{COP}}^{/}=\frac{\left(h_1-h_4^{/}\right)}{\left(h_2^{/}-h_1\right)}& \left(9\right)\left[\mathrm{from}\mathrm{eq}\left(3\right)\mathrm{and}\left(4\right)\right]\end{array}$

Equations (8) and (9) clearly depict the increase in Coefficient of performance obtained by the cooling unit 300 of FIG. 4. For the same refrigerant and the same value its mass flow rate in the refrigeration cycle, the cooling unit 300 obtains a substantially enhanced cooling capacity with a considerably lower value of the power required to drive the compressor 420, as compared to the conventional cooling unit of FIG. 3. Additionally, the problem of condensate removal from the cooling units based on vapor-compression refrigeration cycles is significantly alleviated by the inventive cooling unit 300.

FIG. 7 further illustrates the heat transfer phenomenon within the condenser of the cooling unit 300 of FIG. 4 in accordance with the present disclosure. As shown, control volume 700 surrounds the condenser 430 and all the heat exchange processes at the condenser side occur within the control volume 700. Since the condenser 430 is air-cooled, hot outdoor air 710 acquires an average temperature T₁ after the condensate 492 is sprinkled into the airstream, enters the control volume 700 and leaves the condenser 430 at a higher average temperature T_m after gaining heat from the refrigerant. Refrigerant 730 enters the condenser 430 at a pressure P₂^/ and temperature T₂^/ (corresponding to the state 2^/ marked in FIG. 6) and emerges at the same pressure but a lower temperature T₃^/ after rejecting heat to the hot air 710 blown into the condenser 430. The temperature values T₁ and T_m for the hot air entering and exiting the control volume 700 are measurable through any suitable temperature detecting device known in the art, and conventionally used for the purpose, including a thermocouple, a thermostat etc. Neglecting the heat gained by the interiors of the condenser 430, including the coil tubes through which the refrigerant 730 flows, and the minute heat losses to the surroundings, the major part of the heat lost by the refrigerant 730 goes into the hot air 710 entering the condenser 430, thus resulting in an elevation in its temperature from an average inlet temperature value T₁ to the outlet value T_m.

• • Let {dot over (m)}_air=mass flow rate of the air blown into the condenser;
  • {dot over (m)}_ref=mass flow rate of the refrigerant through the vapor-compression cycle for the cooling unit 300
  • C=average specific heat capacity of the hot air within temperature range T₁ to T_m
  • h₂^/, h₃^/=enthalpies of the refrigerant at entry point 2^/ and exit point 3^/, as marked in the pressure-enthalpy diagram of FIG. 6
  • q_ref=average rate of heat lost by the refrigerant 730 within the control volume 700
  • q_air=average rate of heat gained by the hot air 710 within the control volume 700
  • We have the following approximations under above mentioned considerations:

q _ref ={dot over (m)} _ref(h₂^/−h₃^/)   (10)

q _air ={dot over (m)} _air C(T_m−T₁)   (11)

• • Neglecting the minor losses within the condenser:

q_ref=q_air

{dot over (m)} _ref(h₂^/−h₃^/)={dot over (m)}_airC(T_m −T₁) [By equating eqns. (10) and (11)]

(h₂^/−h₃^/)={dot over (m)}_airC(T_m−T₁)/{dot over (m)}_ref   (12)

This calculated value of the change in enthalpy of the refrigerant during its flow through the condenser is used as a reference for approximating the extent of the point 2^/ (shown in FIG. 6) until which the refrigerant needs to be compressed by the compressor 420 of the cooling unit 300.

For a specific temperature (T₁) of the air 710 at the inlet to the condenser 700, the value of the higher pressure P₂^/ to which the refrigerant should undergo compression in the compressor 420, for maximizing the cooling capacity of the cooling unit, is obtainable by the standard pressure-enthalpy chart for the refrigerant used. With known values of the mass flow rate of the refrigerant ({dot over (m)}_ref) and the air ({dot over (m)}_air), and an approximate value of the specific heat capacity of the air in the temperature range T₁ to T_m, the corresponding difference between the enthalpies of the superheated vapor refrigerant (entering the condenser) and the saturated liquid refrigerant (exiting the condenser), i.e., (h₂^/−h₃^/), is calculated from Eq. (12).

The high pressure value of the refrigerant (P₂^/) is calibrated to be in conformity with the enthalpy difference calculated in eq. (xii) using the standard pressure-enthalpy chart for the refrigerant used, and this calibrated value P₂^/ is set for the compressor 420 of the cooling unit 300. As an example, using R-134 (a) as the working refrigerant fluid, with the high pressure side saturated vapor temperature of 50° C. (considering hot summer outdoor air entering the condenser), the enthalpies of the saturated liquid refrigerant and the saturated vapor refrigerant are found to be 270 KJ/Kg. K and 420 KJ/Kg. K respectively, from the standard pressure-enthalpy chart for R-134 (a). Considering specific operating values of the mass flow rate of the hot air and the refrigerant in Eq. (12), if the enthalpy difference h₂^/−hd 3^/ is 200 KJ/Kg K (for instance), then the refrigerant needs to be compressed to a pressure of about 1.4 Mpa (P₂^/), and to a temperature of about 90° C. (T₂^/), for proper functioning of the cooling unit 300.

Although the present invention has been described in considerable details with reference to certain preferred versions thereof, other versions are also possible.

The cooling unit as disclosed herein can be used in several circumstances where a refrigerating effect is desired. In an aspect, the cooling unit can be an integral part of a usual air-conditioning systems utilized in homes or other buildings. As another example, several such cooling units can be simultaneously used in collaboration for commercial and industrial applications where large scale air-conditioning is required, including residential buildings, factories etc. As a further example, the unit can also be used in conditioning the air in movie theatres, concert halls, restaurants, cafeteria etc. The appropriate method of use would be to install the evaporator of the cooling unit at a suitable location within the space where the refrigerating effect is desired such that it can extract heat from the space, condition the air and reject this heat elsewhere. These and other variations are well within the scope of those of ordinary skill in the art.

Claims

1. A cooling unit employing vapor compression refrigeration cycle for cooling an ambient medium, the cooling unit comprising: an evaporator, including a plurality of heat exchange tubes adapted to carry a primary heat exchange medium, and arranged to receive a flow of an ambient medium;a condensate pump positioned to route condensate from the evaporator;a sub-cooling heat exchanger in fluid communication with the condensate pump; anda plurality of spray nozzles in simultaneous fluid communication with the sub-cooling heat exchanger and positioned to spray condensate into the flow of the ambient medium.

2. The cooling unit of claim 1, wherein the sub-cooling heat exchanger has a condensate inlet for a continuous inflow of the condensate therein, and a condensate outlet for a continuous outflow of the condensate therefrom.

3. The cooling unit of claim 2, wherein the condensate outlet has a cross-sectional area preferably smaller than the cross-sectional area of the condensate inlet.

4. The cooling unit of claim 2, wherein the condensate inlet is positioned at a higher elevation with respect to the condensate outlet.

5. The cooling unit of claim 2, wherein the sub-cooling heat exchanger is in fluid communication with a condensate outlet pipe, the condensate outlet pipe having the spray nozzles mounted.thereon.

6. The cooling unit of claim 5, wherein the condensate outlet pipe is provided with a set of perforations to fluidly communicate with the plurality of spray nozzles.

7. The cooling unit of claim 1, wherein the sub-cooling heat exchanger has a primary heat exchange medium inlet for a continuous inflow of the primary heat exchange medium therein, and a primary heat exchange medium outlet for a continuous outflow of the primary heat exchange medium therefrom.

8. The cooling unit of claim 1, wherein the primary heat exchange medium and the condensate are in thermal communication within the sub-cooling heat exchanger.

9. The cooling unit of claim 1, wherein the sub-cooling heat exchanger is in fluid communication with least one of the plurality of spray nozzles.

10. The cooling unit of claim 1, wherein the plurality of spray nozzles are arranged in a spaced array.

11. The cooling unit of claim 1, wherein the plurality of spray nozzles are coupled to a condenser, and configured to continuously spray the condensate on air blown into the condenser.

12. A method of increasing the coefficient of performance of a cooling unit utilizing vapor compression refrigeration cycle, the cooling unit including an evaporator, a compressor, a condenser and an expansion valve, the method comprising: collecting condensate from the evaporator;routing condensate to a sub-cooling heat exchanger;exchanging heat in the sub-cooling heat exchanger between the condensate and a primary heat exchange medium;directing the condensate from the sub-cooling heat exchanger to a plurality of spray nozzles; andsprinkling the condensate through the plurality of spray nozzles over air blown into the condenser.

13. The method of claim 12 wherein the cooling unit includes a compressor, and a condenser configured to receive a flow of an ambient medium, the method further comprising compressing the primary heat exchange medium within the compressor to a high-pressure value based at least on a set of parameters corresponding to the ambient medium.

14. The method of claim 13, wherein the set of parameters includes a mass-flow rate value of the ambient medium flowing through the condenser.

15. The method of claim 13, wherein the set of parameters include the corresponding average temperature values of the ambient medium entering and leaving the condenser.

16. The method of claim 12, wherein the condensate remains in continuous thermal communication with a primary heat exchange medium within the sub-cooling heat exchanger.

17. The method of claim 12 further comprising directing the condensate through a condensate outlet pipe from the sub-cooling heat exchanger to the spray nozzles.

18. The method of claim 17 further comprising providing a set of perforations on the condensate outlet pipe to enable it being in simultaneous thermal communication with the plurality of spray nozzles.