JP2017519182A - Improved evaporative condenser - Google Patents

Abstract

An evaporative condenser (10) for use in a cooling or air conditioning system comprises one or more condensing coils (12) arranged in a condensing coil zone (13). The coil condenses system refrigerant within the coil. The condenser also includes a mechanism (14, 15) for wetting one or more condensing coils (12). The condenser further comprises a drift eliminator (30) arranged to remove free water from the air stream A flowing through the one or more condensing coils and the wetting mechanism (14, 15). The condenser is additionally a diverging zone (40) that diverges from the condensing coil zone (13) towards the drift eliminator (30), when the airflow flows through one or more condensing coils (12); A divergence zone (40) is provided in which the air flow is adapted to flow through the divergence zone 40 to the drift eliminator (30).

Description

Improved evaporative condensers and evaporative condensation processes for use in cooling and air conditioning systems are disclosed. The condenser and process may be used with both chemical refrigerants (eg, hydrofluorocarbons) and natural refrigerants (eg, hydrocarbons (eg, propane and isobutane), CO ₂ , ammonia, etc.).

  Existing evaporative condensers are used in various cooling and air conditioning systems to exhaust heat due to refrigerant condensation. More particularly, the evaporative condenser comprises one or more wetted (eg sprayed) condensation coils. The condensing coil is for condensing the refrigerant when an air flow passes through the condensing coil, and heat is removed from the refrigerant in the condensing coil by evaporating a part of the water, and the refrigerant in the condensing coil To condense. The evaporative condenser also includes a drift eliminator (or, more simply, an eliminator, where “drift” is water that will pass to the atmosphere if not removed). The drift eliminator removes free water that passes with the air flow as the air flow flows through the condensing coil and water spray. Free water removal is done before releasing the air stream to the atmosphere.

  In existing evaporative condensers, the planar area of the condensing coil is commensurate with the planar area of the drift eliminator, ensuring a constant air flow rate and air flow velocity through the evaporative condenser.

  In existing evaporative condensers, the heat exchange efficiency is limited by the speed of air flowing over the condensation coil. The speed of the air is now limited by the ability of the drift eliminator to remove free water from the air passing through the drift eliminator. In existing evaporative condensers, the water thus removed is recycled for reuse when wetting the condensation coil. However, any water that passes through the drift eliminator with the air flowing to the atmosphere may contain bacteria, such as Legionella, so that as much free water as possible is required to be removed from the air stream.

  For example, many existing evaporative condensers are known to regulate the maximum air velocity through the drift eliminator to 3.5-4 m / s to ensure sufficient water removal. At high maximum air velocities, it is still assumed that there is a significant risk of bacteria (eg Legionella) passing through the drift eliminator with free water that has not been removed. A safer maximum air velocity of 3.5 m / s through the drift eliminator has been proposed. However, this in turn will set a limit on the velocity of air that can flow over the condensing coil.

  Reference to the background art above does not necessarily constitute recognition that the technology forms part of the common general knowledge of those skilled in the art. Also, the above references are not intended to limit the applications of the condensers and processes disclosed herein.

Disclosed herein is an evaporative condenser for use in a cooling or air conditioning system. The evaporative condenser disclosed herein includes chemical refrigerants (eg, hydrofluorocarbons, hydrochlorofluorocarbons, perfluorocarbons, hydrofluoroolefins, etc.) and natural refrigerants (eg, hydrocarbons, eg, propane and isobutane, CO ₂ , ammonia, etc. ) Can be condensed.

  The evaporative condenser disclosed herein comprises one or more condensing coils, which condense the system refrigerant within the condensing coils. One or more condensing coils may be located in the condensing coil zone of the evaporative condenser. The condensing coil zone may comprise an air plenum having a constant cross-sectional area.

  The evaporative condenser disclosed herein also includes a mechanism for wetting one or more condensation coils (eg, by spraying with water).

  The evaporative condenser disclosed herein further comprises a drift eliminator arranged to remove free water from the air stream flowing through the one or more condensing coils and the wetting mechanism.

  According to the present disclosure, the evaporative condenser disclosed herein comprises a divergence zone that diverges from the condensation coil zone toward the drift eliminator. The configuration of the diverging zone is such that when the air flow flows through one or more condensing coils, the air flow flows through the diverging zone to the drift eliminator. For example, the divergence zone may comprise an air plenum having a gradually increasing cross-sectional area.

  The divergence zone can reduce the speed of the air flow exiting the condensation coil zone. This means that the speed of air passing over the condensation coil can be increased relative to the speed of air passing through the drift eliminator. This higher speed can help reduce tube fouling.

  Furthermore, it has been surprisingly found that a condensing coil bundle having a reduced planar area relative to the drift eliminator can be used. Furthermore, this requires fewer condensing coils for the same condenser performance. This means that a lower cost evaporative condenser can be produced. This is because the condenser coil bundle represents the only most expensive component of such a condenser.

  In addition, the increased refrigerant flow may pass through the condensing coil bundle. This is because higher air velocities can cause a relatively large amount of refrigerant to condense.

  In addition, as an alternative to using known hot dip galvanized carbon steel condensation tubes, more expensive and / or stronger materials (eg stainless steel) can be used to form one or more condensation coils, resulting in Longer life, less corrosion, and optionally thinner coil (tube) walling can be used. Nevertheless, if preferred, DN 8, 10, 15, and 20 (schedule 40) seamless, hot-dip galvanized carbon steel tubes may still be used to form one or more condensation coils. .

  In one embodiment, each of the one or more condensing coils is a stainless steel tube (eg, 304 or 316 stainless steel, with an outer diameter of 4.76-31.8 mm and a thickness of 0.5-1.6 mm. May be used. The use of 304 stainless steel can provide better conductivity, while 316 stainless steel can provide better corrosion resistance. Such a tube material can work favorably compared to known condensation coil tubes of galvanized mild carbon steel. The use of very small diameter tubes may be suitable for certain small scale applications.

The use of stainless steel tube material (ie, due to corrosion / chemical resistance, increased refrigerant pressure strength, etc.) is also acceptable using natural refrigerants such as propane and / or isobutane hydrocarbons, CO ₂ , ammonia, etc. .

  In one embodiment, one or more condensing coils may be arranged in the condensing coil zone as a bundle (eg, a bundle of two or more nested coils). For example, the condensing coil zone may comprise a section of the condenser with a generally constant cross-sectional area (eg, an air plenum with a hollow section such as circular, square, rectangular, etc.).

  In one embodiment, the divergence part of the zone may be configured to gradually decelerate the air flow.

  In one embodiment, the diverging zone may comprise a hollow frustum (a hollow air plenum) through which the air flow flows. Such a hollow frustum may be located on the air outflow side of the condensing coil plenum. For example, if the condensing coil plenum is a circular section, each diverging frustum includes a frustum, or a square to circular frustum-shaped prism, and if the condensing coil plenum is a quadrilateral section, the diverging frustum is a rectangular frustum. May be included.

  In one embodiment, the drift eliminator may be located on the air outflow side of the divergence zone.

  In one embodiment, the condenser may comprise an intake chamber located on the air inflow side of the condensation coil zone.

  In one embodiment, the mechanism for wetting one or more condensing coils may comprise one or more spray nozzles. The spray nozzle may be positioned relative to the divergence zone to spray water onto the one or more condensing coils in a direction opposite to the air flow flowing through the one or more condensing coils. For example, the spray nozzle may be located in the divergence zone, and water may be sprayed into the condensing coil zone as a generally liquid cone.

  Alternatively, the mechanism for wetting one or more condensing coils may comprise a water distribution channel, such as a water distribution channel having serrated edges, internal slots, and the like.

  The condenser may comprise a water recovery zone (eg located at the bottom of the intake chamber). The recovery zone can recover the water that has passed through the condensation coil zone.

  The condenser may further comprise a recycling system for wetting the collected water and recycling it to the mechanism to maximize the efficiency of the condenser. In one embodiment, the recycling system may include a pump for wetting the collected water through the tubing and feeding it to the mechanism. For example, an extraction pipe may extend from the bottom of the intake chamber to the pump, and a delivery pipe may extend from the pump outlet to the wetting mechanism (eg, to a spray nozzle, water pipe, etc.).

  In one embodiment, the recycling system may further comprise a water replenishment mechanism to maintain a predetermined amount of water (eg, in a water recovery zone) for effective operation of the evaporative condenser, if desired. . Such makeup water may include water removed (trapped) by a drift eliminator.

  In one embodiment, the evaporative condenser may further comprise a heat exchanger (eg, a separate, externally located separate heat exchange unit). The recovered water may pass through a heat exchanger before being recycled to the wetting mechanism. In addition, the condensed refrigerant may pass through a heat exchanger to exchange heat with recovered water that is recycled. Such a heat exchanger may be used to subcool the condensed refrigerant and further improve the operating efficiency of the evaporative condenser.

  The present specification discloses an evaporative condenser with a recovery zone for recovering water that has passed through the condensing coil zone, is equipped with a heat exchanger, and the recovered water is heated before being recycled to the wetting mechanism. The condensed refrigerant that passes through the exchanger passes through the heat exchanger and exchanges heat with the recovered water that is recycled.

  Also disclosed herein is an evaporative condensation process that forms part of a cooling or air conditioning cycle.

  The process includes passing the refrigerant through one or more condensation coils. The process also includes the step of wetting one or more condensation coils with water. The process further includes condensing the refrigerant in the coil by passing the air stream over one or more wet condensing coils and evaporating a portion of the water into the air stream. The process additionally includes removing water present in the air stream exiting the one or more condensing coils.

  According to the present disclosure, the process is performed to reduce the velocity of the air flow exiting the one or more condensing coils before removing the water present in the air flow.

  As outlined above, this allows the planar area (and hence the amount) of one or more condenser coils to be reduced relative to the drift eliminator (along with the attendant advantages outlined above).

  Also disclosed herein is an evaporative condensation process where water passing through one or more condensing coils is collected and recycled to wet one or more condensing coils with water. Further, in such a process, heat may be exchanged between the condensed refrigerant and the recovered water before the recovered water is recycled to wet one or more condensing coils.

  The process disclosed herein may occur in the evaporative condenser described above.

In the processes disclosed herein, the refrigerant that condenses in one or more condensing coils can be a natural refrigerant (eg, hydrocarbons such as propane and / or isobutane, CO ₂ , ammonia, etc.) or a chemical refrigerant (eg, hydrofluorocarbon, hydro Chlorofluorocarbon, perfluorocarbon, hydrofluoroolefin, etc.).

It is a figure which shows the outline of the cross-section side surface of the evaporative condenser which has the condensing coil zone by which the 1 or more condensing coil is arrange | positioned, and the divergence zone extended in the direction away from a condensing coil zone. It is a figure which shows the detail of FIG. 1, and shows the modification of the evaporative condenser which further contains the heat exchanger of a side surface. FIG. 3A is a schematic diagram showing a cross-section of an evaporative condenser having a convergence-divergence zone in which one or more condensation coils are arranged, and FIG. 3B shows a convergence-divergence zone in which one or more condensation coils are arranged. It is a figure which shows the outline of the side surface of the evaporation condenser which has. FIG. 2 is a schematic cross-sectional side view of an evaporative condenser similar to FIG. 1, but for different process parameters, according to an embodiment. ₂ is a graph showing the temperature profile of CO ₂ and water according to an example. 4 is a graph showing a CO ₂ heat capacity profile according to an example. It is a graph which shows the water flow which flows down a tube bundle by an Example. It is a graph which shows an overall heat-transfer coefficient and pressure loss by an Example. FIG. 4 is a graph showing an exhaust heat profile based on a commercially available transcritical CO ₂ compressor at a saturated suction temperature of 5 ° C., a useful suction superheat of 5K, and a CO ₂ liquid temperature of 5 ° C. according to an example. According to the embodiment, it is a graph showing the performance of a commercial 50Hz of transcritical CO ₂ compressor, 30kW / 27.2m ³ / h. According to the _embodiment, it is a graph showing the variation due to NH 3, R22, R507A, saturated condensing temperature of propane and R134a the COP.

  Regardless of the condenser and any other form that may belong to the scope of the process described in the Summary of the Invention, specific embodiments will be described by way of example only with reference to the accompanying drawings. Specific forms of evaporative condensers and evaporative condensation processes that form part of a cooling or air conditioning system / cycle are described.

  The evaporative condenser embodiments designated 10 and 100 are shown in FIGS. 1 and 2 and FIGS. 3A and 3B, respectively. Evaporative condenser embodiments 10 and 100 may use both chemical and natural refrigerants (described above). 4 to 11 relate to the embodiment described in the examples.

  1-3, similar components of evaporative condensers 10 and 100 have similar numbers, but 100 has been added to the embodiment of FIG. Furthermore, for the sake of brevity, in the following description, these similar or similar components that reappear in the embodiment of FIG. 3 have not been described again, so they are to be interpreted as already described. It should be understood that it should.

  The preferred evaporative condenser 10 of FIGS. 1 and 2 comprises two or more nested condensing coil bundles 12 within which the selected refrigerant of the system flows (for condensation). . The condensation coil bundle 12 is arranged in a condensation coil zone in the form of a rectangular airflow plenum 13.

The evaporative condenser 10 also includes a mechanism in the form of a spray nozzle 14 formed in the distribution tube 15, which can be achieved by spraying the condensing coil bundle 12 with a cone of water 16 (eg, as shown). , Speed 3 kg / m ² ), for wetting the condensing coil bundle 12. Alternatively, a water distribution channel, such as a water distribution channel with serrated edges or internal slots may be used.

  The spray nozzle 14 is arranged to spray water onto the condensing coil bundle 12 in the opposite direction of the air flow flowing through the condensing coil bundle 12 as shown.

  The evaporative condenser 10 also includes a fan disposed in the fan housing at the upper end of the condenser. Such an arrangement is actually shown in the embodiment of FIG. 3 as a fan 118 located in a fan housing 120 located at the top end of the condenser (see FIG. 3A). The same or similar arrangement can be used in the embodiments of FIGS. In this regard, the fan draws air into the intake chamber 22 from the intake port 21, and the intake chamber 22 is disposed toward the lower end of the condenser 10.

In the embodiment of FIGS. 1 and 2, the air flow A enters, for example, at a volumetric flow rate of 8.1 m ³ / s, first passes through the mesh filter, then enters the intake chamber 22 and is then condensed by the fan. It flows upward through the coil bundle 12. The air pressure difference may be maintained at 160 Pa, for example, by a fan.

  In the embodiment of FIG. 3, the air stream A enters at a velocity of, for example, 3 m / s and a wet bulb temperature of, for example, 23 ° C., and first passes through an optional mesh filter 124 and intake slot 126 in response to air contamination. , Enters the intake chamber 122, and then flows upward through the condensing coil bundle 112 by the fan 118.

  The evaporative condenser 10 further includes a drift eliminator 30, and the drift eliminator 30 is disposed in the condenser adjacent to the upper end of the condenser. The drift eliminator 30 removes free water from the air stream as free water flows through the condensing coil bundle 12 and the spray nozzle 14.

  In the embodiment of FIGS. 1 and 2, the evaporative condenser 10 comprises a rectangular airflow plenum 13 immediately followed by a diverging airflow zone in the form of a frustum-shaped plenum 40. The rectangular airflow plenum 13 may be a square, rectangular, etc. hollow section plenum (eg, a folded or joined plastic or metal sheet / plate). The diverging plenum 40 may also be a hollow section (eg, a plastic or metal sheet / plate joined by folding) but is formed to define a frustum. For example, if the plenum 13 is a square section plenum, the diverging plenum 40 comprises a square or rectangular frustum.

  However, in the embodiment of FIG. 3, the evaporative condenser 100 uses both a converging airflow zone 135 and a diverging airflow zone 140, which are located on either side of an intermediate rectangular airflow plenum 113. Yes. An intermediate rectangular airflow plenum 113 houses the condensing coil bundle 112. The plenum 113 has a constant cross-sectional area and interconnects the convergent air flow zone 135 and the diverging air flow zone 140. The intermediate airflow plenum 113 may again be a square, rectangular or other hollow section (eg, sheet / plate) plenum. The converging air flow zone 135 and the diverging air flow zone 140 may again be hollow sections (eg, sheets / plates), but are each formed to define a frustum. For example, if the intermediate zone 113 is a square section zone, the converging and diverging frustums may each comprise a square or rectangular frustum.

  In the embodiment of FIGS. 1 and 2, the fan is operated such that the airflow A is already at a higher speed in the condensing coil bundle 12 than the drift eliminator 30. As the air stream A flows through the condensing coil bundle 12, it flows into the diverging air stream plenum 40 and passes through the water cone 16. As the cross-sectional area of the diverging air flow plenum 40 gradually increases, the air flow can be decelerated to an acceptable speed before reaching and passing through the drift eliminator 30. The evaporative condenser 10, particularly the diverging air flow plenum 40, is configured so that this speed is at a level where free water in the minimum environmentally acceptable amount of air flow can be removed. In this regard, the air flow rate in the drift eliminator 30 can be reduced to about 3.5 m / s.

  It will be appreciated that because the drift eliminator 30 is located in the immediate vicinity of the air outlet of the diverging air flow plenum 40, the air flow is not allowed to decelerate more than necessary.

  Accordingly, the embodiment of FIGS. 1 and 2 does not use a convergent airflow zone. Rather, the air flow velocity from the intake chamber 22 through the condensing coil bundle 12 is about 5 m / s until the air flow reaches the diverging air flow plenum 40, and the air flow is about 3.5 m at the drift eliminator 30. Decelerate gradually until / s.

  However, in the embodiment of FIG. 3, the condensing coil bundle 112 is located in the intermediate airflow zone 113. The configuration of these zones is such that the airflow A flows through the converging airflow zone 135 to the condensing coil bundle 112 located in the intermediate airflow plenum 113, where it is accelerated (e.g. To about 5 m / s). As the air flow A flows through the condensing coil bundle 112, it flows into the diverging air flow zone 140, passes through the water cone 16, and decelerates before reaching the drift eliminator 130. Again, the drift eliminator 130 is located in the immediate vicinity of the air outlet of the diverging air flow plenum 140.

  The convergent air flow zone 135 is configured to gradually accelerate the air flow A, for example. Conversely, the diverging airflow zone 140 is configured to gradually decelerate the airflow, for example. This is because the velocity of the air passing over the condensing coil bundle 112 through the intermediate air flow zone 113 is relative to the velocity of the air entering the intake chamber 122 and to the velocity of the air through the drift eliminator 130. Means increase. For example, in the depicted configuration, the air flow rate in the intermediate air flow zone 113 is about 5 m / s (ie, about 45% higher) that is about twice the velocity of air 3.5 m / s through the drift eliminator. ).

  In either embodiment, as a result of the increased air flow rate over the condensing coil bundles 12, 112, surprisingly, condensing coil bundles having a reduced planar area relative to the drift eliminators 30, 130 can be used. Was discovered. Furthermore, it has been discovered that as a result of this increased air flow, fewer condensing coils are required for the same condenser heat rejection performance.

As a result, a lower cost evaporative condenser can be manufactured. This is because the condensing coil bundle represents the only most expensive component of the condenser. Alternatively, instead of using a known thick-walled hot dip galvanized carbon steel condensation tube in the coil bundles 12, 112, the coil bundles 12, 112 are formed using a more expensive and / or stronger material, such as stainless steel. May be. In such cases, a longer coil life, less corrosion, and, if desired, a wall material for thinner coil bundle tubes. In this regard, the coil bundles 12, 112 are stainless steel tubes, such as 304 or 316 stainless steel, with an outer diameter of 4.76-31.8 mm and a thickness of 0.5-1.6 mm. May be. Such a tube is observed to work well compared to the known condensation coil tube of galvanized mild carbon steel. Corrosion and chemical resistance, as well as increased refrigerant compressive strength, are provided by such stainless steel tube materials, but such increased characteristics are due to natural refrigerants such as propane and / or isobutane hydrocarbons, CO ₂ , Ammonia or the like can be used in the evaporative condenser 10,100.

  Another consequence of the increased air flow over the condensing coil is that an increased refrigerant flow can pass through the condensing coil bundles 12,112. This is because higher air velocities can cause a relatively large amount of refrigerant to condense.

  The condenser 10 also includes a water recovery zone in the form of a basin 50 that is located at the bottom of the intake chamber 22 (ie, adjacent to the intake chamber 22). The basin 50 collects excess spray water that has passed through the condensing coil.

  In order to maximize the efficiency of the condenser, the condenser 10 further comprises a recycling system for recycling the recovered water to a distribution tube 15 for supply to the spray nozzle 14. In this regard, the recycling system includes a pump 52 for sending the collected water to the distribution tube 15 through the piping. The pump 52 draws water from the basin 50 by the take-out pipe 54. At this time, the delivery pipe section 56 extends from the pump outlet and connects to the distribution tube 15.

  The recycling system also includes a water replenisher 58 (for example, 383 kg / h) for maintaining a predetermined amount of water in the basin 50 for the evaporative condenser to operate effectively. Such make-up water may include a supply of water removed (trapped) by the drift eliminator 30.

  In the variant of the evaporative condenser shown in detail in FIG. 2, the condenser 10 may further comprise a side heat exchanger unit 60. The water in the basin 50 may be pumped into the heat exchange unit 60 by the pump 52 before being recycled to the distribution tube 15 by the delivery pipe section 56. Such a unit may also be suitable for the embodiment of FIG.

In this variation, the refrigerant condensed in the condenser tube may also pass through the heat exchange unit 60 via the refrigerant delivery pipe 62 to exchange heat with the water recycled from the basin 50. In the heat exchange unit 60, the refrigerant in which the water in the relatively low temperature basin is condensed can be subcooled to, for example, around 30 ° C to 26.5 ° C. This can further improve the operating efficiency of the cooling system. The refrigerant (for example, CO ₂ ) flowing out as the flow 64 through the heat exchange unit 60 is at a subcooled temperature (for example, around 26.5 ° C.).

  Provide a non-limiting example of the present condenser and process, provide a theoretical basis for the condenser and process, and better understand the condenser and process during operation.

[Example 1-Process design model]
Design model for application to subcritical CO ₂ condensation of evaporative condenser, developed a design model depicted in Figures 1-3, for example. More specifically, the benefits of applying evaporative condensation techniques to subcritical CO ₂ condensation were investigated. Such benefits included lower design pressure, lower energy consumption, and lower running and operating costs compared to trans-critical operation. It has also been noted that hot gas defrosting can be a standard feature of subcritical CO ₂ cooling plant operation.

First, however, it was noted that ammonia can condense at 30 ° C. in an evaporative condenser with an inlet air wet bulb temperature of 24 ° C. The developed design model showed that an evaporative condenser for condensing subcritical CO ₂ at 30 ° C. (ie, 1.1 K below the critical point) can be designed with a wet bulb temperature of 24 ° C.

Secondly, many European average climatic conditions, including the warmer climates of Spain, Italy, Greece and Turkey, are suitable for evaporative condensers and note that subcritical CO ₂ evaporates at 30 ° C. It was done. It was noted that Canada, the wide range of the USA, China, and most of Australia below the southern regression line also had a climate suitable for applying evaporative condensers to subcritical CO ₂ condensation. It was noted that the thermodynamic and transport properties of subcritical CO ₂ at 30 ° C. changed significantly with temperature. Thus, in a specific design, we also showed the effect these changes have on the CO ₂ temperature profile, heat transfer, and pressure loss.

For example, according to a survey of average climatic conditions, many of Europe, including Spain, Italy, Greece, and Turkey, have condensation temperatures below 30 ° C. and always in many places to condense CO ₂ in a subcritical state. It became clear that the evaporative condenser has a climate that can be applied. For example, the only place in Europe where the design wet bulb temperature incidence of 5% exceeds 24 ° C was Adana in Turkey (wet bulb incidence design levels of 1% and 2.5%) Is 26 ° C.). In Thessaloniki, Greece, the 1% wet bulb design incidence is 25 ° C, while the 2.5% and 5% wet bulb design incidence levels are 24 ° C. The next highest 1% wet bulb design incidence level, 24 ° C, occurred in Gibraltar, Barcelona, Valencia, Milan, Istanbul and Izmir.

Finally, the CO _2, the use of evaporative condenser in temperate climates and many subtropical climate, may be used CO ₂ cooled everywhere to the same extent as any chemical refrigerant, and should be used in an indirect application It was concluded that if there are (eg office buildings and hospital heating and cooling), it can compete well with ammonia.

About 20 years ago, when CO ₂ cooling became popular again, air cooled gas cooling (partially adiabatic assistance by spraying water on the intake surface of a finned coil gas cooler) ) Has been applied almost universally. As a result, it was noted that virtually all CO ₂ cooling systems need to operate in transcritical mode. This is because the air cooling temperature is close to or exceeds the critical temperature of CO ₂ , 31.1 ° C.

Mostly, the summer design CO ₂ outlet temperature from the air cooled gas cooler was higher than the critical temperature, and as a result, the compressor had to operate at a pressure of 90 bar or higher to ensure a reasonable COP. . The summer design COP of a transcritical CO ₂ compressor is generally lower than an air cooled HFC or evaporatively cooled ammonia system.

Accordingly, it has been proposed to reduce the temperature of the condenser cooling medium to a level that allows for a complete subcritical CO ₂ cooling cycle. This is achieved with an evaporative condenser in which the ambient air wet bulb (WB) temperature is an effective coolant temperature, where the ambient air dry bulb (in the case of an air cooled condenser or gas cooler) DB: Dry Bulb) temperature was not used.

  Notable issues include water supply, water consumption and water treatment needs, and control of minimum condensation temperature, which is currently mandated by some compressor suppliers. Another problem was the control strategy for handling unintended transcritical states. Recommendations have been made to address these issues.

[Evaluation model for CO ₂ evaporation condenser]
[Evaluation example]
FIG. 4 shows a schematic flow sheet of an evaporative condenser that will now be further described. In the flow sheet of FIG. 4, water was recycled on the tube bank and the spray water temperature was the same as the basin water temperature.

The parameters specified are: (a) air velocity and wet bulb temperature and dry bulb temperature, (b) spray water flow rate, (c) bundle size, and (d) effluent CO containing saturated liquid at 30 ° C. and 7.2 MPa. ₂ .

[Mass and energy balance of evaporative cooler]
Qureshi (2006) and Heyns (2009) have published five simultaneous nonlinear differential equations that explain air-fluid interactions in evaporative cooling.

  The equation was solved by writing the program using a fourth-order Runge-Kutta routine written in Microsoft VBA behind the Microsoft Excel spreadsheet, dividing the path length into 40 intervals. The solution was trial and error. This is because the basin water temperature at the air inlet was repeatedly estimated and adjusted until it was the same as the calculated water temperature at the air outlet.

The solution proceeded “reversely” from the CO ₂ outlet to the inlet along the tube path. The saturated liquid refrigerant started from the inlet at 30 ° C., proceeded upwards as if to heat, and ended with superheated steam at the calculated exhaust temperature. The program allowed both two-phase condensation and single-phase steam desuperheating.

[Model validation]
There was no analytical solution for the five equations that could validate the numerical solution. However, the following two findings were noted: if the effluent and influent temperatures were equal, (a) the CO ₂ enthalpy change was equal to the wet air enthalpy change and (b) calculated for ammonia condensation at 30 ° C. The heat load was within 9% of the load calculated using the Merkel model based on a simplified constant condensation temperature (Merkel, 1926).

[Model prediction]
5 and 6 show the temperature profile of CO ₂ and water. The shape of the CO ₂ temperature profile (FIG. 5) was surprising and was much flatter than expected. Over approximately 37% of the exchanger surface, the CO ₂ vapor temperature only decreased from 32 ° C. to 30 ° C. in interval number 29. This was the result of a very high heat capacity (Figure 6) just above 30 ° C.

The model predicted that 67% of the exchanger surface would be required for sensible cooling. For CO ₂ , enthalpy data at 30 ° C. near the critical point showed that 68% of the exhaust heat was sensible heat cooling. This is different from ammonia where sensible heat cooling is only 10%.

The water temperature profile was distorted to the left and reflected a larger proportion of sensible heat cooling for CO ₂ near the critical point compared to a profile with relatively small sensible heat cooling.

[Evaporation of water]
FIG. 7 shows the water flow flowing down the tube bundle. Another surprising thing was that water did not evaporate into the air at the top of the bundle. Here, the water temperature is low but rising, and since the water is in contact with air at an absolute humidity higher than the absolute humidity at the air-water interface, some condensation occurs and the water flow increases. .

[Influence of characteristic changes]
FIG. 8 illustrates the effect of changes in heat capacity, density, viscosity, and thermal conductivity on the overall heat transfer coefficient and pressure loss over the respective solution intervals. The zero interval was where hot exhaust gases entered. The overall heat transfer coefficient increased considerably as the CO ₂ temperature approached 32 ° C., and correspondingly the pressure loss per meter decreased as the vapor phase transitioned to two phases.

  In the model, 0.845 was used for the Lewis number in equation (3). At the Lewis number of 1.00, the surface area required for the same heat load was reduced by only 1.4%.

[opinion]
The modeled case was extreme in that the condensation of CO ₂ at 30 ° C. was very close to the critical point. It was noted that at lower condensing temperatures, the rate of sensible cooling would decrease and the variation in properties with temperature would be greatly reduced. Reference was made to FIG.

It is further noted that the exhaust heat predicted by the simplified Merkel model was about 22% lower than the differential model of CO ₂ condensation at 30 ° C., given the significant rate of sensible cooling, It was not unexpected.

[CO ₂ compressor subcritical energy performance]
[Effect of condensation temperature on cycle performance]
In FIG. 10, five COP plots were generated for a commercially available semi-sealed transcritical CO ₂ compressor with a 4-pole motor with a stroke volume of 27.2 m ³ / h, 50 Hz, and 30 kW.

  Referring to curve 1 in FIG. 10, at a saturated suction temperature (SST) of + 10 ° C., the COP is from 6.27 (at a saturated condensation temperature (SCT) of + 30 ° C.) to 18.0. (In SCT at + 16 ° C.). An SST of + 10 ° C. would allow an evaporation temperature (ET) of + 11 ° C. with a suction pressure drop corresponding to a boiling point drop of 1K. 11 ° C. is noted as a reasonably efficient evaporation temperature for direct cooling of AC (Air Conditioning) air, and allows a relatively large diffusion of air temperature across the cooling coil, The volume of air that would need to be circulated was limited, thereby reducing fan energy consumption and the resulting parasitic heat load. This in turn will increase the overall energy efficiency of the entire system by leading to a reduction in the required energy input to the compressor.

  Curve 2 showed a COP ranging from 4.45 to 11.67 at an SST of + 5 ° C. and an SCT of 30 ° C. to 16 ° C. This will allow additional cold water production for AC to be added to existing buildings and applied to new buildings.

In both of the above two cases, the AC compressor could also act as a parallel compressor for the cooling load at -5 ° C SST. Such a cooling load is, for example, a high stage load for a two-stage CO ₂ system that is applied to maintain a refrigerated storage temperature near 0 ° C. and for cold storage and blast freezing applications.

In such a case, the high stage compressor will operate at virtual CO ₂ gas cooler outlet temperatures + 5 ° C. and + 10 ° C. This results in COP curves 3 and 4, respectively. The COP curve 3 was in the range of 4.7 to 7.88 with a virtual gas cooler outlet temperature of + 5 ° C. with SST of −5 ° C. and SCT ranging from + 30 ° C. to + 16 ° C. COP curve 4 showed a COP in the range of 4.45 to 7.04 at an SST of −30 ° C. to + 16 ° C. with an SST of −5 ° C. at a virtual gas cooler outlet temperature of + 10 ° C. It has been noted that this can be improved with the suction heat exchanger (SHEX) of the compressor suction and can result in performance closer to curve 3.

[Influence of ambient wet bulb temperature on condenser performance]
FIG. 4 shows the overall details for CO ₂ , air, and water. The effect of ambient wet bulb temperature on condenser performance is shown in the results described in the table below.

[Relative energy efficiency of ammonia, R22, R507A, propane, and R134a]
The COPs of these refrigerants under the same operating conditions are shown in FIG. From the results, it was confirmed that ammonia was the best refrigerant among these refrigerants. Surprisingly, the COP of R134a was low. In SCT at 16 ° C. and 35 ° C., the COP of R134a was 42% and 31% lower than that of ammonia, respectively. Furthermore, the COP of the R134a compressor in the SCT at + 16 ° C. was 3.84, which was almost the same as the COP of the ammonia compressor in the SCT at + 35 ° C. under the same suction conditions. Thereby, it was confirmed that R134a has both a high direct and indirect global warming potential (GWP). The performance of R507A was 11% to 16% less efficient than R22 in SCT from 25 ° C to 35 ° C. HFC R507A does not have Ozone Depletion Potential like HCFC R22, but R507A's 100-year GWP is 3,895, more than twice the R22's 100-year GWP, 1,810 high.

[Overall heat transfer factor, Uo]
Referring again to FIG. 8, Uo was much higher than the familiar Uo for ammonia condensation. For ammonia condensation, Uo is about 450 to 550 w / m ² at a surface air velocity of 2.6 to 3.05 m / s. K range. A surface air velocity of 3 m / s was chosen as the maximum value of the model, ensuring that the drift eliminator could capture most of the free water floating in the upper air draft.

For the CO ₂ evaporative condenser of FIG. 4, the average Uo of FIG. 8 is about 1,050 w / m ² . K. It was noted that this exceeded twice the average value for ammonia at virtually the same surface air velocity entering the condenser tube bundle. This is, as shown in FIG. 9, with 30 ° C. condensation, 68% of the heat removed is sensible superheat, and only 32% is actually the latent heat of condensation at 30 ° C. Given that, it was notable. The high overall heat transfer factor is a high CO ₂ mass flux of 338.7 kg / m ^{2 in a} 76 off 55 meter equivalent length circuit causing a calculated pressure drop of 15 kPa. s. This is the maximum value that can be tolerated to facilitate the CO ₂ condensers operating in parallel and avoiding liquid hold-up without requiring too large drop legs.

Similar to the evaporator, the high ΔP / ΔT ratio of CO ₂ allowed a high mass flux in the condenser circuit. This also helped make tube bundles more economical because they gave higher rates of heat transfer and allowed for fewer and longer circuits.

The ammonia mass flux of the evaporative condenser is about 25-40 kg / m ² . s range, often 25 kg / m ² . It was noted that it was lower than s. The pressure drop was a concern in the ammonia condenser. This is because an excessive pressure drop in the ammonia evaporative condenser raises the discharge pressure, which also raises the saturation condensation temperature (SCT), resulting in increased energy consumption.

[Minimum air flow results]
Referring back to FIG. 4, the calculated outflow air dry bulb temperature was 29.3 ° C. at 100% RH, so the outflow wet bulb temperature was also 29.3 ° C. This was 0.7 ° K lower than the 30 ° C SCT. This was possible because the upper tube was at a temperature of 77 ° C. and a high rate of sensible heat overheating ensured a high spill approach TD of 47.7 ° K. available. It was noted that this was not possible with an ammonia evaporative condenser. In an ammonia evaporative condenser, the minimum effluent temperature approach between the ammonia SCT and the effluent wet bulb temperature was rarely less than 3K, and was 2.5K or higher in design conditions. Also, with less airflow, the fan's energy consumption is minimized.

[Conclusion]
Applying an evaporative condenser with a maximum design wet bulb temperature (WB) of 24-25 ° C in a higher latitude subtropical region, subject to obtaining a satisfactory performance test of the full scale prototype CO ₂ evaporative condenser It was concluded that it was very promising. The CO ₂ evaporative condenser was even more promising in temperate climates with lower ambient WB temperatures, and in cold to cold climate areas.

Based on the above conclusions, areas suitable for applying evaporative condensers to subcritical CO ₂ compressor exhaust gas condensation include virtually all of Europe (including the Mediterranean countries), the United States (facing the Gulf of Mexico and the Atlantic Ocean). In most southern states, and most of the midwestern states located to the north of Minnesota).

Experiments have also shown that evaporative gas cooling with an ambient wet bulb temperature of 28-29 ° C. and an ambient air WB-CO ₂ outlet temperature approach of 3K is completely feasible. This is because in the transcritical mode, there is no condensed phase (FIG. 9), there is only a larger LMTD sensible heat transfer, a relatively high transcritical fluid density and a high heat capacity similar to that shown in FIG. It was due to the fact that there was.

Evaporative cooling was applied to both subcritical CO ₂ condensation and transcritical CO ₂ gas cooling, leading to efficient cooling of high COPs. This high COP was comparable to that achieved with conventional refrigerants operating below the critical point and was often higher than that achieved with conventional refrigerants. This has opened the way for applying CO ₂ cooled the worldwide. This, CO ₂ is in the air-conditioning load at the compressor saturated suction temperature of + 5 ° C. and + 10 ° C., applications used for cold water, as well as the DX or pump CO 2 _AC applications respectively, especially true.

  It was further noted that the AC compressor could also act as a parallel compressor for any remaining cooling load in the facility. The facility is, for example, a supermarket and is a place where refrigeration and refrigeration loads are required with high to very high COP as shown in FIG.

In fact, comparing FIGS. 10 and 11, as found also by Pearson (2010), in the subcritical condensed phase, CO ₂ performs better than conventional chemical refrigerants such as R22, R507A, and R134a. Was clearly superior. In addition, CO ₂ was comparable to or superior to ammonia and propane, under most operating conditions, especially when including parallel compression.

For example, at a high wet bulb temperature of 28 ° C., a conventional evaporative condenser can be operated with an SCT of 40 ° C., and as shown in FIG. 9, the COP is NH ₃ , R 22, R 507 A, propane, and R 134 a, respectively. It would have been 3.37, 3.34, 2.71, 2.96, and 2.38. Therefore, a larger compressor, such as a modified CNG fuel compressor, is needed to develop high efficiency CO ₂ cooling.

[Terminology]
In the examples, it is as follows.

[Model parameters]
1. NIST (2011) data was used for the thermodynamic and transport properties of saturated and superheated CO ₂ .
2. The h _w of equation (6) was calculated from equation (A.8) of Mizushima and Miyashita (1967), Qureshi and Zubair (2006).
3. _{H d} of the formula (3) is, of Mizushima and Miyasita (1967), was calculated from the equation (A.13) of Qureshi and Zubair (2006).
4. For _two- phase CO ₂ flow, the h _i in equation (6) is calculated from equations (A.6) and (A.7) in Shah's (2009), Qureshi and Zubair (2006), and the pressure loss Was calculated from Muller-Steinhagen and Heck correction (ASHRAE, 2005).
5. For single-phase CO ₂ vapor flow, _{h i} of Equation (6) computes the Dittus-Boelter correction Nu ^{= 0.023Re} 0.8 ^{Pr 0.3,} pressure loss, friction coefficient = 0.079Re ^-0. Calculated from ²⁵ .
6). The decrease in air pressure across the tube bank is described in Mills (1999), section 4.5.1, p. Calculated from 316.

The following references were used to create the model.
1. ASHRAE, 2005, 2005 Fundamentals, page 4.12-13
2. Heyns J, Kroger D, 2009, Performance characteristics of an air-cooled steam condenser inducing a hybrid (dry / wet) dephlegmator, AppendixAp-AP
3. Merkel, F.M. , 1926, Verdung skunking, VDI-Zeitschrift, Vol. 70, pp. 123-128
4). Mills A. F. 1999, Basic Heat & Mass Transfer, 2nd ed. A. F. , Prentice Hall.
5. Mizushima, T .; , R. Ito and H.M. Miyashita, 1967, Experimental study of an evolutionary cooler, International Chemical Engineering, Vol. 7, pp. 727-732
6). NIST 2011, http: // webbook. nist. gov / chemistry / fluid /, Thermophysical Properties of Fluid Systems
7). Qureshi B, Zubair S, 2006, A complete design and rating study of efficient coolers and condensers. Part I Performance evaluation, Int. J. et al. Refrigeration, 29: 645-658.
8). Shah M, 2009, An improved and extended general correlation for heat transfer condensation in plain tubes, HVAC & R Research, 15 (5)
9. Pearson, S.M. Forbes, 2010, Use of carbon dioxide for air conditioning and general refrigeration, IIR-IOR 1st Cold Chain Conference, Cambridge, UK.

[Example 2-Design model output]
In order to show the variation of condenser capacity with surface air velocity, the following data points were created by the design model.

  Although several condenser and process embodiments and models have been described, it should be understood that the condenser and process may be embodied in many other forms.

  For example, the plenum 13 may be a circular section, in which case the diverging plenum 40 includes a truncated cone or a square to circular frustum-shaped prism. However, such a configuration is not very preferable. This is because free drainage of water in the condenser is not promoted.

  In the claims below and in the foregoing description, the word “comprise” and variations thereof, such as “comprises” or “ “Comprising” is used in an inclusive sense, ie, to define the presence of the described feature, but in various embodiments of the condensers and processes disclosed herein, Used to not exclude the presence or addition of additional features.

Claims (19)

An evaporative condenser for use in a cooling or air conditioning system,
One or more condensing coils arranged in the condensing coil zone, the condensing coils for condensing the refrigerant of the system in the condensing coils;
-A mechanism for wetting the one or more condensing coils;
A drift eliminator arranged to remove free water from the air stream flowing through the one or more condensing coils and a wetting mechanism;
A divergence zone that diverges from the condensing coil zone towards the drift eliminator, wherein the air flow flows through the one or more condensing coils, the air flow passes through the divergence zone and the drift eliminator. A divergence zone that flows into
An evaporative condenser.   The condenser according to claim 1, wherein the one or more condensation coils are arranged in the condensation coil zone as a bundle.   The condenser of claim 2, wherein the condensing coil zone comprises a section of the condenser having a generally constant cross-sectional area.   The condenser according to any one of claims 1 to 3, wherein the diverging zone is configured to decelerate the airflow flowing through the diverging zone before reaching the drift eliminator.   The condenser according to any one of claims 1 to 4, wherein the diverging zone comprises a hollow frustum through which the air flow flows.   The condenser according to any one of claims 1 to 5, wherein the drift eliminator is located in the immediate vicinity of the air outlet side of the divergence zone.   The condenser according to claim 1, further comprising an intake chamber located on an air inflow side of the condensation coil zone.   The mechanism for wetting the one or more condensing coils comprises a spray nozzle, wherein the spray nozzle causes water to flow in a direction opposite to the air flow flowing through the one or more condensing coils. The condenser according to claim 1, which is arranged with respect to the divergence zone so as to spray.   The condenser of claim 8, wherein the nozzle is disposed in the divergence zone for spraying the water onto the one or more condensation coils as a generally liquid cone.   The condenser according to any one of claims 1 to 9, further comprising a recovery zone for recovering water that has passed through the condensing coil zone, and a recycling system for recycling the recovered water to the wetting mechanism.   The recycling system maintains a pump for pumping the collected water through the piping to the wetting mechanism and, if necessary, a predetermined amount of water for the evaporative condenser to operate effectively. The condenser according to claim 10, further comprising a water replenishment mechanism.   A heat exchanger, wherein the recovered water passes through the heat exchanger before being recycled to the wetting mechanism, and the condensed refrigerant passes through the heat exchanger and is recovered for recycling. The condenser according to claim 10 or 11, further comprising a heat exchanger for exchanging heat with water.   The condenser according to any one of the preceding claims, wherein each of the one or more condensing coils comprises a stainless steel tube. An evaporative condenser for use in a cooling or air conditioning system,
One or more condensing coils arranged in the condensing coil zone, the condensing coils for condensing the refrigerant of the system in the condensing coils;
-A mechanism for wetting the one or more condensing coils;
A drift eliminator arranged to remove free water from the air stream flowing through said one or more condensing coils and a wetting mechanism;
-A recovery zone for recovering water that has passed through the condensing coil zone;
-A recycling system for recycling the collected water to the wetting mechanism;
-A heat exchanger, wherein the recovered water passes through the heat exchanger before being recycled to the wetting mechanism and the condensed refrigerant passes through the heat exchanger and is recycled. A heat exchanger that exchanges heat with the treated water,
An evaporative condenser.   15. A condenser according to claim 14, as defined in any one of claims 1-13. An evaporative condensation process that forms part of a cooling or air conditioning cycle,
-Passing the refrigerant through one or more condensing coils;
-Wetting the one or more condensing coils with water;
-Condensing refrigerant in the coil by passing the air stream over the wet one or more condensing coils and evaporating a portion of the water in the air stream;
-Removing water present in the air stream flowing out of the one or more condensing coils;
Including
An evaporative condensation process in which the velocity of the air stream exiting the one or more condensing coils is decelerated before removing the water present in the air stream. An evaporative condensation process that forms part of a cooling or air conditioning cycle,
-Passing the refrigerant through one or more condensing coils;
-Wetting the one or more condensing coils with water;
-Recovering and recycling the water passing through the one or more condensing coils to wet the one or more condensing coils with water;
-Condensing refrigerant in the coil by passing the air stream over the wet one or more condensing coils and evaporating a portion of the water in the air stream;
-Removing water present in the air stream flowing out of the one or more condensing coils;
-Exchanging heat between the condensed refrigerant and the recovered water before recycling the recovered water and wetting the one or more condensing coils;
Including evaporative condensation process.   The process according to claim 16 or 17, wherein the process takes place in an evaporative condenser according to any one of claims 1-15.   19. The process of any one of claims 16-18, wherein the refrigerant that condenses in the one or more condensing coils comprises a chemical refrigerant or a natural refrigerant, e.g., a refrigerant described herein.