METHOD AND APPARATUS FOR A DELAYED AND PROLONGED AIR COOLING CONDENSATION SYSTEM

Abstract

In various embodiments devices and methods are provided for an improved dry-cooling condensation system. In certain embodiments the methods involve receiving steam from a source of steam (e.g., a power plant); condensing the steam into water while transferring the latent heat of the steam into the latent heat of a thermal storage material; and dissipating the latent heat from the thermal storage material at a later time when the ambient temperature is lower than the ambient temperature at the time the steam was condensed into water.

Description

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims priority to and benefit of U.S. Ser. No. 61/533,551, filed Sep. 12, 2011, which is incorporated herein by reference in its entirety for all purposes.

STATEMENT OF GOVERNMENTAL SUPPORT

[Not Applicable]

BACKGROUND

In electrical power plants, large amounts of latent heat carried by low temperature steam from the steam turbine exhausts needs to be removed and condensed into water to complete the Rankine cycle for electricity generation. Similar situations exist in other heat engines and appliances, such as central air-conditioning systems. Typically, the low quality (temperature) latent heat was removed by the use of a large amount of cooling water, which is a method used by the majority of coal fired power and nuclear power plants. However, this so called “wet-cooling” method consumes extremely large amounts of water, accounting for almost 50% of total industrial water usage and 20% of total water consumption.

For water resource scarce areas, “dry-cooling” methods were developed to directly dissipate the latent heat into ambient air without using a lot of water. There are two major types of air-cooling condensation methods, one is direct air-cooling condensation and the other is indirect air-cooling condensation. These two methods are distinguished from each other by whether the heat transfer process occurs between low temperature steam and ambient air directly or through other heat transfer media.

The heat transfer efficiency of air-cooling condensation strongly depends on the temperature difference between the exhaust steam and the environment ambient air, and the speed of the air that blows onto the condensation heat exchanger surfaces. To dissipate the latent heat of exhausted steam instantaneously requires a larger temperature difference between the exhausted steam and the ambient air, typically as much as 30° C. to 40° C. in “dry-cooling” methods, and 10° C. to 15° C. in “wet-cooling” methods. Moreover, since the volume heat capacity of the air is almost a thousand times smaller than that of the water, an extremely large amount of cooling air is blown upon the heat exchanger surfaces for the cooling. The large temperature difference between the exhaust steam and the ambient air results in higher exhaust steam pressure. For example, in a typical dry-cooling system, the exhaust steam pressure is typically about 25 to 50 kPa instead of 5 to 7 kPa for a typical wet-cooling system. This exhaust steam pressure increase results in a drop in heat to electricity conversion efficiency for the steam turbine generator. For example, the heat to electricity conversion efficiency would decrease from 40% to 35% when “dry-cooling” is used instead of a “wet-cooling” for a typical 600 MW coal-fire power generator.

On the other hand, during a “dry-cooling” process, the heat dissipation rate depends on the surface area of the condenser's heat exchanger and the wind blowing speed that passes through the heat exchanger surfaces. Due to cost considerations, the heat exchanger's surfaces cannot be too large otherwise the wind speed has to be very large to achieve effective cooling. According to the aerodynamics, the power consumption of a wind-blowing machine used for dry-cooling varies as the 3rd power of the wind speed it generates. The higher the wind speed it requires, the more electricity it consumes. Therefore, the dry-cooling approach may consume a few percent up to more than 10 percent of the electricity that the same power plant generates.

SUMMARY

To overcome above mentioned shortcomings of an air-cooling condensation process, it was discovered that it is beneficial to use an indirect air-cooling approach while dividing the cooling process into two time periods or “cycles”: 1) A condensation period where the low temperature exhaust steam transfers its latent heat to a low temperature thermal storage material while being condensed into water, e.g., during a time period when ambient air temperature is relatively high; and 2) A dissipation period where the heat is dissipated from the thermal storage material to the ambient air by air-cooling, e.g., during a time period when ambient air temperature is relatively low. If these two stages are separated in time it is possible to reduce the shortcomings of the above mentioned dry-cooling methods. This two stage (time period) method can also be applied partially, e.g., in certain embodiments, during the high temperature period, only a part of the generated latent heat is stored in thermal storage materials, and dissipated into ambient air during the cool time period; while other part is still dissipated using conventional cooling methods. This is especially useful to modify and improve existing power plants.

It is even more beneficial for a concentration solar thermal power (CSP) plant to use a dry-cooling method because in most cases, CSP power plants are constructed in desert areas where water resources are very limited. In this case, a separated two-stage dry-cooling can bring even more advantages compared to the traditional dry-cooling methods because the temperature differences between day and night are normally larger in desert areas than in other locations.

The apparatus and methods described herein have been developed in view of the above points. Accordingly, in various embodiments, methods and apparatus are described that provide delayed and prolonged dry-cooling condensation methods, that take the advantage of different ambient air temperatures during the day and night to lower the energy consumption for cooling (e.g., for a wind blower) while decreasing the condensation temperature of the exhaust steam so that the overall thermal to electricity conversion efficiency can be improved for dry-cooling.

Let the starting time of the first condensation stage to be t₁, and the duration of the stage 1 to be Δt₁; the starting time of the second heat dissipation stage from the thermal storage media to be t₂, and the duration of stage 2 to be Δt₂. If t₁ equals t₂ and Δt₁ equals to Δt₂, the entire cooling process becomes a normal indirect air-cooling condensation process. For a delayed and prolonged indirect air-cooling condensation, t₂ (the starting time of the dissipation stage) is typically later than t₁ (e.g., at a cooler time of a given day), and in certain embodiments, Δt₂ is not equal to Δt₁. In this way, effectiveness of the condensation process can be optimized independently from the subsequent heat dissipation process, thereby improving the effectiveness of the air-cooling process.

In certain embodiments by taking advantage of the fact that the ambient air temperature during the night is lower than that during the day, one can use low temperature thermal storage materials to store the latent heat of exhausted steam during the day while condensing the turbine exhaust steam into water, and then dissipate the stored thermal energy into the ambient air during the night when the ambient temperature is lower, using a wind blower to drive the ambient air through the packaged thermal storage material surface to release the latent heat stored therein into the environment. In certain embodiments low temperature phase change materials (PCMs) are used to store the latent heat from the exhaust steam (during Δt₁ time period). The stored thermal energy is then dissipated into the ambient air during the night when the ambient air temperature is lower than temperature during the day via wind blower in a pre-determined time duration Δt₂.

In addition, when circumstances require it, one can choose time durations for the condensation stage Δt₁ and the heat dissipation stage Δt₂ so that the heat dissipation rate can be slower than that of a typical indirect dry-cooling system. In certain embodiments one can chose a time duration Δt₂>Δt₁. Thus, the exhaust steam condensation temperature can be significantly decreased. It should be noted that for every 10° C. of higher condensation temperature for a steam turbine generator, the heat to electricity conversion efficiency would be lower by 4.5%. This is a very significant loss of efficiency because normally a “dry-cooling” approach raises the condensation temperature up to 30° C. higher than that of a “wet-cooling” approach.

In certain embodiments if the delay and prolonged indirect dry-cooling approach described herein is used, in the stage of condensation for the exhausted steam from the steam turbine or other heat appliances, the low temperature exhausted steam transfers its latent heat to the low temperature thermal storage media via an efficient heat exchanger while condensing the steam into water. The exhaust steam condensation temperature (typically a few degrees higher than the low temperature thermal storage material's phase change temperature) can be set at only a few degrees higher than the ambient air temperature at night. During the heat dissipation stage, the phase change material would have enough temperature difference relative to the ambient air temperature so that the stored latent heat can be dissipated into the environment, as described in detail below. This is how the new approach described herein can lead to an improvement for the heat to electricity conversion efficiency.

As described in detail below, the large amount of heat exchange area of our apparatus described herein ensures that the steam can be condensed into water at close to phase change temperature during the condensation stage, and that the stored latent heat can be dissipated into the environment during the night, resulting in a drop of wind speed required for effective cooling. This can reduce the electricity consumption significantly due to the third power relationship between the power needed to drive a wind blower and the wind speed that is created by that blower.

Accordingly, in various embodiments, device for steam condensation and delayed dissipation of the heat produced by the condensation is provided. In certain embodiments the device comprises a combined condensation/thermal storage chamber, the chamber comprising: one or more valved ports for receiving steam from a steam source; one or more containers containing a thermal storage material; one or more valved ports for applying a vacuum to the thermal storage chamber; one or more valved ports for introducing ambient pressure air into the chamber; one or more valved ports for removing condensed water from the chamber; and a valve system operably coupling the chamber to a source of ambient temperature air. In this context a valved port indicates that flow of fluid or gas through the port is under the control of one or more valves, however, the valves need not be located at the site of the port and can be remote (e.g., disposed between a vacuum source and a vacuum port, between a steam source and a steam port, and the like). In certain embodiments the one or more containers containing a thermal storage material is a plurality of containers each containing a thermal storage material. In certain embodiments the thermal storage material is a phase change thermal storage material (PCM). In certain embodiments the thermal storage material is a liquid/solid phase change thermal storage material. In certain embodiments the thermal storage material comprises a material selected from the thermal storage materials shown in Table 2 (e.g., Na₂CO₃.10H₂O, and the like). In certain embodiments the phase change material contains glass microfibers or nanofibers. In certain embodiments the device further comprises an apparatus to cause mixing of liquid phase thermal storage material in the containers containing the thermal storage material. In certain embodiments the containers (containing the thermal storage material(s)) are attached to a frame structure (see, e.g., FIG. 3) and the apparatus comprises a motor configured to cause rotation of the structure and the attached containers. In certain embodiments the one or more valved ports for applying a vacuum to the thermal storage chamber are operably coupled to a vacuum pump. In certain embodiments the one or more valved ports for applying a vacuum to the thermal storage chamber and the one or more valved ports for introducing ambient pressure air into the chamber are controlled by separate valves. In certain embodiments the one or more valved ports for applying a vacuum to the thermal storage chamber and the one or more valved ports for introducing ambient pressure air into the chamber are controlled by the same valve(s). In certain embodiments the valve system operably coupling the chamber to a source of ambient temperature air comprises a butterfly valve (or a flange cover, or other functionally equivalent structure). In certain embodiments the source of ambient air is a fan and/or blower and/or one or more ducts configured to receive ambient wind. In certain embodiments the device is one of a plurality of the devices configured in a parallel configuration. In certain embodiments the one or more valved ports for receiving steam from a steam source are operably coupled to the low temperature steam output from a turbine (e.g., a turbine in a power plant selected from the group consisting of a coal-fired power plant, a gas-fired power plant, a nuclear power plant, and a solar thermal power plant). In certain embodiments the device is sited in a desert or a non-desert region with limited water availability.

Also provided are systems for delayed heat dissipation from the condensation of waste steam, said system comprising a plurality of devices for steam condensation and delayed dissipation of the heat produced by the condensation, e.g., as described herein. In certain embodiments the devices are configured in a parallel configuration, e.g., as illustrated in FIG. 2. In certain embodiments the system comprises at least 10, or at least 20, or at least 30, or at least 40, or at least 50, or at least 60, or at least 70, or at least 80, or at least 90, or at least 100, or at least 150, or at least 200 of said devices. In certain embodiments the system is operably coupled to the low temperature steam output from a turbine. In certain embodiments the turbine is a turbine in a power plant (e.g., a coal-fired power plant, a gas-fired power plant, a nuclear power plant, a solar thermal power plant, etc.). In certain embodiments the system is sited in a desert or a region with limited water availability.

In various embodiments a dry-cooling methods are provided. It will be recognized that the methods, as described herein provide methods of condensing waste steam while delaying heat dissipation from said condensation; and can be used to complete a Rankine cycle at enhanced efficiency. Accordingly, in certain embodiments a dry-cooling condensation method is provided where the method comprises: receiving steam from a source of steam; condensing the steam into water while transferring the latent heat of said steam into the latent heat of a thermal storage material; and dissipating the latent heat from said thermal storage material at a later time when the ambient temperature is lower than the ambient temperature at the time the steam was condensed into water. In certain embodiments the thermal storage material is a phase change thermal storage material (PCM). In certain embodiments the thermal storage material is a liquid/solid phase change thermal storage material. In certain embodiments the thermal storage material comprises a material selected from the thermal storage materials shown in Table 2 (e.g., Na₂CO₃.10H₂O, and the like). In certain embodiments the source of steam is the steam output from a turbine. In certain embodiments the turbine is in a power plant selected from the group consisting of a coal-fired power plant, a gas-fired power plant, a nuclear power plant, and a solar thermal power plant. In certain embodiments the receiving and condensing comprises receiving and condensing during daylight hours. In certain embodiments the receiving and condensing comprises receiving and condensing peak temperature hours (e.g., between noon and 3:00 pm). In certain embodiments the dissipating comprises dissipating the latent heat from the thermal storage material during cooler hours (e.g., the late afternoon, and/or evening, and/or night). In certain embodiments the method is performed using one or more devices for steam condensation and delayed dissipation of the heat produced by the condensation, e.g., as described herein. In certain embodiments the receiving and condensing comprises: opening said one or more valved ports for applying a vacuum to a thermal storage chamber to reduce the ambient pressure in the thermal storage chamber; opening one or more valved ports for receiving steam to introduce steam from a steam source (e.g., a turbine) into the chamber, whereby the steam condenses transferring latent heat of steam into a thermal storage material; and operating one or more valved ports for removing condensed water from the chamber to return the condensed water to the system providing the steam. In certain embodiments the dissipating comprises: restoring the pressure in the thermal storage chamber to atmospheric pressure; operating a valve system operably coupling the chamber to a source of ambient temperature air to pass ambient temperature air through the thermal storage chamber to transfer heat from the thermal storage material to the air. In certain embodiments the passing ambient temperature air comprising operating a fan and/or blower or to force air through the chamber, and/or coupling the chamber to a duct system that channels wind through the chamber. In certain embodiments the dissipating further comprises operating an apparatus to provide mixing of fluid thermal storage material in chambers containing the fluid thermal storage material. In certain embodiments the method comprises operating a motor to rotate a structure frame to which the chambers containing the thermal storage material are attached. In certain embodiments the method is performed using a system comprising a plurality of In certain embodiments the method is performed using one or more devices for steam condensation and delayed dissipation of the heat produced by the condensation, e.g., as described herein. In certain embodiments the devices are configured in a parallel configuration. In certain embodiments substantially all of the devices in said system perform said receiving and condensing at the same time. In certain embodiments substantially all of the devices in said system perform said dissipating at the same time. In certain embodiments some of the devices perform said receiving and condensing at the same time that other devices are dissipating.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a schematic diagram of the apparatus 100 operably linked to a steam source (e.g., a steam turbine 101). An efficient condensing heat exchanger is used to convert the latent heat of exhausted steam (e.g., from a power plant) into the latent heat of a thermal storage medium while the exhausted steam is condensed into water during a “condensation stage”. During a second “heat dissipation stage”, the valves (e.g., butterfly valves, or flange covers) are opened and the stored heat in thermal storage medium is dissipated into the ambient air by an air stream (wind) by a blower or other wind source. The labeled parts are described as follows: steam source (e.g., turbine) 101; steam exhaust valve 102 that connects the steam turbine and the condensing heat exchanger/thermal storage tank; vacuum pump 103 that is used to pump the air inside the tank out to create a vacuum (reduced pressure) in the tank; a vacuum valve 104 that connects the vacuum pump and the tank; micro pole 105 for the vacuum pathway; valves 106 (e.g., butterfly valves, flange covers, or other functionally comparable structure(s)) that seal the vacuum during the “condensation stage”; wind blower 107 (or other source of an ambient temperature air stream); a thermal storage tank 108; a supporting base 109 for the thermal storage tank; packaging containers 110 (e.g., pipes) for the thermal storage materials (e.g., phase change materials (PCMs); supporting structure 111 for the phase change material packaging pipes (e.g., as illustrated in FIG. 3) that, in certain embodiments, is supported by an rotates on bearing 117; condensed water passage way 112 from the tank; shutoff valve 112A for use during a “heat dissipation stage”; an apparatus (e.g., a motor) 113 that rotates the entire phase change packaging system (e.g., comprising a structure frame 115 attached to the packaging containers containing the thermal storage material), e.g., via a helical drive system 116 to mix the phase change thermal storage materials inside the thermal storage packaging pipe to prevent the possible phase separation inside the pipes; air release valve 114 to vent the air into the tank until the valve 106 can be opened at the beginning of the “heat dissipation stage”.

FIG. 2 schematically illustrates another embodiment of the device, where multiple steam condensation tank/heat dissipation devices (e.g., as illustrated in FIG. 1) are configured in a parallel system/structure 200 operably coupled to a steam source 201. The labeled parts are described as follows: steam source (e.g., turbine) 201; N number of similar modules 202 as described in FIG. 1 (illustrated with N=5); wind path pipes 207 that connect the condensing tanks together to the source of the air stream; the wind blower 204 (or other source of a cooling air stream); the valve 205 (e.g., a butterfly valve, flange cover or other functionally equivalent structure) that allows air from the blower to enter the thermal storage chamber 108 (when operating as a heat dissipation apparatus); valve 206 (e.g., a butterfly valve, flange cover or other functionally equivalent structure) that allows the air to blow out of the heat dissipation apparatus; the parallel wind path pipes 203 to allow the air stream to flow out of system; valve 208 that connects the steam source (e.g., steam turbine exhaust) with the condensation/heat dissipation system. Vacuum pump 213 connects (e.g., is operably coupled) via vacuum valve 214 and pipe 215 connects with the main exit pipe 203. Its function is to reduce pressure in the entire parallel condensation/heat dissipation system before the turbine exhaust steam is released into the system for condensation.

FIG. 3 illustrates a structure frame 115 in a cross section view for the packaged thermal storage pipe fixture and arrangement. 110 is the cross section of the packaged thermal storage pipes; 116 is the structure frame to hold the storage pipes in the storage tank. The center distance between packaged pipes can be optimized to allow the air stream to pass through the pipes along the length direction of the storage tank.

DETAILED DESCRIPTION

In various embodiments devices are described that provide delayed and prolonged dry-cooling condensation methods that take advantage of different ambient air temperatures during the day and night to reduce energy consumption for cooling (e.g., for operating a wind blower in a dry-cooling system) while decreasing the condensation temperature of the exhaust steam so that the thermal to electricity conversion efficiency can be improved for dry-cooling.

Methods are described in here in detail with reference to devices examples of which are illustrated in the accompanying drawings. The methods and devices are intended to be illustrative and not limiting. Using the teaching provided herein variations of the illustrated devices and methods will be readily available to one of skill in the art.

Aspects of the innovations, such as those set forth in some of the implementations below, may relate to systems and methods of air-cooling. However, it should be understood that the inventions herein are not limited to any such specific illustrations, but are defined by the scope of the claims and full disclosure.

As illustrated in FIG. 1 a vacuum can be applied to the chamber 108 by running a vacuum pump 103 (or other source of vacuum) after opening of vacuum valve 104 of the thermal storage tank 108 via a micro pole 105. Low temperature exhaust steam from a steam source 101 (e.g., from a power plant turbine) enters thermal storage tank 108 via opened valve 102, to release the latent heat of the stem to containers 110 (e.g., packaged pipes) holding thermal storage materials (e.g., phase change thermal storage materials (PCMs)). This transfer of latent heat from the steam to the thermal storage material is accompanied by condensation of the steam into water. The water drains out from the tank 108 via water path (e.g., optionally valved ports) 112 to complete the Rankin cycle for the power (e.g., electricity) generation. Where the thermal storage material is a phase change material, the phase change material (PCM) thermal storage material inside containers (e.g. pipes) 110 absorbs the latent heat of steam and changes its phase from solid form into liquid form, which means the phase change material (PCM) melts and keeps (stores) the latent heat. This process comprises the “condensation stage”.

The maximum temperature difference required for the exhaust steam and the phase change temperature of the thermal storage material can be described by the following equation:

$T_w-T_m=\frac{{\rho }_{\mathrm{cs}}\frac{d_o^2}{4}{\Phi }_1^2\Delta t_1}{2{\gamma }_c{{\lambda }_{\mathrm{cs}}\left({\rho }_{\mathrm{cs}}V_c\right)}^2}=\frac{{\rho }_{\mathrm{cs}}{\gamma }_cd_o^2}{8{\lambda }_{\mathrm{cs}}\Delta t_1}$ $\mathrm{where}$ $V_c=\frac{Q_1}{{\rho }_c{\gamma }_c}=\frac{\Delta t_1{\Phi }_1}{{\rho }_c{\gamma }_c}$

where T_w is the wall temperature for the packaging pipe, which is also very close to the exhaust steam temperature for the thermal storage material, T_m is the phase change temperature, ρ_cs (ρ_c) is the phase change material density in the solid state, γ_c is the heat of fusion of the phase change material (or latent heat during phase change), λ_cs is the thermal conducting coefficient for the solid state phase change material, V_c is the total volume; d_o is the diameter for the phase change material's packaging container (e.g., pipe), Φ₁ is the amount of latent heat for the turbine exhaust steam that needs to be condensed per hour; as defined before, and Δt₁ is the time duration for the “condensation stage”, and Q₁ is the total latent heat condensed during the condensation stage. One can find that if the time duration for the “condensation stage” Δt₁ is larger, the maximum temperature difference is smaller because longer time is used to transfer the steam latent heat to the PCMs.

The following illustrates the use of this equation for a real world example: Consider small scale CSP power plant with 10 MW_e peak capacity. With the help of the delay and prolonged dry-cooling apparatus described herein, the exhaust steam condensation temperature for the steam turbine can be 35° C. or below. In this case, its heat to electricity conversion efficiency would be 35%. Therefore, about 65% of the total thermal energy produced by a solar thermal field needs to be eventually condensed into ambient air during power generation. Assume such a CSP system would work six hours to produce electricity per day. The amount of latent heat for the turbine exhaust would be 15 MW_th. To apply the delay and prolonged dry-cooling approach described herein 26 individual storage/dissipation units, similar to those illustrated in FIG. 1 can be used to complete the task. In one illustrative embodiment, each storage unit comprises a cylindrical container about 5 meters in diameter and 16 meters in length, made from fiber-glass material with a steel enhancement frame structure. Inside each container, 5200×3 thin glass tubes (with 25 mm diameter and 0.5 mm wall thickness, 5 meter in length with 3 identical sections to fill the 16 meter long tank, as illustrated in FIG. 1) are placed in parallel with the long dimension of the container in a supporting a structure such as the one illustrated in FIG. 3, where the frame structure is designed to hold the thin glass tubes in such a way that the distance between glass tube centers is about 62.5 mm. This distance is an optimized value for the storage tank design after considering the density of the storage pipes in the tank and the wind resistance when the wind is blown along the pipes to dissipate the heat from the storage pipe. An illustrative thermal storage material is Na₂CO₃.10H₂O compound, whose phase change temperature is at 32° C. with the latent heat (heat of fusion) of 267 kJ/kg. The total amount of thermal storage material that is packaged inside the glass tubes in the container is 50 ton, with total latent heat capacity about 3.6 MWh. To store all the turbine exhaust steam latent heat of the CSP plant into the phase change material during the “condensation stage”, which is 6 hours in this case, applying the equation above, the temperature difference of T_w and T_m would be 2.6° C. As mentioned before, 26 such storage units configured in parallel, e.g., as illustrated in FIG. 2, would be sufficient to condense all the exhaust steam from the 10 MW CSP plant into water during 6 hours of “condensation stage”, store all the latent heat into these storage tanks, and wait until in the night time to be dissipated into ambient air.

During the “heat dissipation stage”, e.g., during the nighttime or cooler daylight periods, when the ambient air temperature is significantly lower than the phase change temperature of the thermal storage materials in the storage tank, a vent valve 114 is opened to release the air into the tank 108 until the pressure inside the tank reaches atmospheric pressure, the two butterfly valves 106 are opened, and a fan (wind blower) 107 is started to drive the ambient cooling-air flow through the thermal storage package pipe surfaces via the spaces between the package pipes 110. The ambient cooling air-flow carries the latent heat of the phase change material (PCM) from the package pipes 110 surfaces into the environment. After releasing its latent heat the phase change material (PCM) of the thermal storage medium returns to its solid form again.

In certain embodiments, in order to avoid the nucleation of the phase change material (PCM) inside its container (e.g., pipe) an agitator or mixer can be provided to facilitate mixing of the PCM inside the container(s). In certain embodiments the agitator/mixer can be fixed at the bottom of the thermal storage tank 108. In certain embodiments the agitator/mixer comprises a motor 113 that can drive rotation of the heat pipes attached to structure frame 111 via for example, a helical drive 116 when the thermal storage materials are in liquid phase. As a result, this motion effectively mixes the liquid phase change thermal storage materials inside the pipes to avoid the possible nucleation and phase separation of the thermal storage medium. In various embodiments other agitators/agitation systems can be contemplated. Such systems include, for example acoustic/ultrasound agitation, mechanical vibrators (e.g., piezoelectric vibrators), and the like.

An alternative method to avoid the possible nucleation and phase separation of the thermal storage medium is to mix super thin (e.g., 5-50 μm, or 10-30 μm, or 10-20 μm, etc.) glass fibers (e.g., about 0.1 to about 10%, or 0.5% to about 5%, or about 1% in volume) into the phase change material inside the packing containers (e.g., pipes). Other suitable materials will be recognized by one of skill in the art. Typically, any matrix materials with low density to prevent solids from precipitating to the bottom are suitable.

As noted above, the methods and devices described herein are particularly advantageous because of the separation of the “dry-cooling” process into the “condensation stage” and the “heat dissipation stage”. During the “condensation stage”, the turbine exhaust steam enters the storage tank and convert its latent heat (from vapor to water) into thermal storage material's latent heat with phase change process.

However, the heat dissipation process involves entirely different heat exchange mechanism because it utilizes a forced convection process with air to dissipate the heat inside the packaged thermal storage pipes into the environment.

Consider the process of “heat dissipation stage” in the following illustrative example. Taking the temperature variation in a typical day into account, assume that the lowest temperature of that day equals to (a° C.) and the highest temperature equals to (a+b° C.). Furthermore we propose a hypothesis that the temperature variation within one day can be described as a Sine function, and the summit of this curve (the highest temperature) appears at 14:00 o'clock. We can then use the following formula to describe the ambient air temperature:

$\begin{array}{cc}{T}_f=a+b\sin \left[\frac{\pi }{24}\left(t-2\right)\right]& \left(1\right)\end{array}$

Suppose that the starting time of heat dissipation stage t₂ is later than 2:00 pm, so the duration of the heat dissipation stage t₂ is given by:

Δt₂=52−2t₂   (2)

By integration of the temperature variation function over the entire time duration of stage 2, we obtain the average ambient temperature during the heat dissipation stage:

$\begin{array}{cc}\stackrel{\_}{T_f}=a+b\frac{{\sin }^2\left(\frac{\pi }{96}\Delta t_2\right)}{\frac{\pi }{96}\Delta t_2}& \left(3\right)\end{array}$

According to the relevant aerodynamics, the Nusselt number Nu, Reynolds Number Re and the heat exchange coefficient h are related by:

$\begin{array}{cc}h=\mathrm{Nu}\frac{\lambda }{d_e}& \left(4\right)\end{array}$

where Nu is given by:

Nu=0.023×Re^0.8×Pr^0.33

while Reynolds Number Re is in the range of Re=10⁴−1.2×10⁵; d_e is the equivalent diameter of the thermal storage tank cylinder, λ is the thermal conductivity of the air flow, which is 2.63×10⁻² W/(m K) for air at temperature of 25° C. Reynolds Number Re is given by:

$\begin{array}{cc}\mathrm{Re}=\frac{{\mathrm{ud}}_e}{v}& \left(5\right)\end{array}$

where μ is the velocity of the air flow and v is the coefficient of viscosity of air, which is 15.53×10⁻⁶ m²/s for air at temperature of 25° C. Substituting equation (5) into equation (4), we have:

h=4.04u^0.613d_e^−0.382(W/m²K)   (6)

Moreover, the equivalent diameter of cylinder d_e is expressed by:

$\begin{array}{cc}{d}_e=\frac{4A_c}{P}=\frac{D_i^2-{\mathrm{Nd}}_o^2}{D_i+{\mathrm{Nd}}_o}& \left(7\right)\end{array}$

where D_i is the internal diameter of the cylinder tank, while d_o is the external diameter of the N pipes.

If the heat transfer rate Φ₁ by the exhausted steam from turbine and the time duration of “condensation stage” Δt₁ are known, the requirement of the phase change material (PCM) is then determined. The total volume of the phase change material (PCM) is given by:

$\begin{array}{cc}{V}_{\mathrm{PCM}}=\frac{\Delta t_1{\Phi }_1}{{\rho }_{\mathrm{PCM}}{\gamma }_{\mathrm{PCM}}}& \left(8\right)\end{array}$

where ρ_PCM is the density of the phase change material (PCM) in units of kg/m³; and γ_PCM is the latent heat of the phase change material (PCM), in unite of J/kg.

From the fundamental heat transfer equation:

$\begin{array}{cc}\frac{\Delta t_1{\Phi }_1}{\Delta t_2}=hA\Delta T_m& \left(9\right)\end{array}$

where h is the convection heat transfer coefficient between air flow and outer wall of pipes, in units of W/(m² K); A is the heat exchange area for heat dissipation processes, in units of m²; ΔT_m is the average convection heat exchange temperature difference, i.e. temperature difference between the outer diameter of the main tank surface and cooling air temperature, K; Substitute (6) into (9):

$\begin{array}{cc}\frac{\Delta t_1{\Phi }_1}{\Delta t_2}=A{\mathrm{hT}}_m=4.04u^{0.613}d_e^{-0.382}A\Delta T_m& \left(10\right)\end{array}$

When we consider the formula of air specific heat we have:

Φ₂=C_pq_mΔT_a   (11)

where Φ₂ is the heat dissipation power for the proposed apparatus, C_p is the specific heat at constant pressure of the air flow, in units of J/(kg K); q_m is the mass flow rate of the air in units of kg/s, and ΔT_a is the temperature different of the air flow between inlet and outlet of thermal storage tank 108. In addition, it is noted that Δt₁Φ₁=Δt₂Φ₂ which means that equation (10) equals equation (11).

Another temperature difference that can be taken into account is the difference between the temperature of the phase change material (PCM) and the outer wall of package pipe containing the PCM. This temperature difference relates to the properties of the phase change material (PCM), the external diameter of pipe do and the time duration of stage Δt₁ or Δt₂:

$\begin{array}{cc}T_m-T_w=\frac{{\rho }_{\mathrm{PCM}}{\gamma }_{\mathrm{PCM}}d_o^2}{8{\lambda }_{\mathrm{PCM}}\Delta t}& \left(12\right)\end{array}$

In order to keep the temperature difference between the phase change material (PCM) and the outer wall of the pipes suitable small, the pipes (PCM containers) can be selected such that the outer diameter is less than:

$\begin{array}{cc}{d}_o\le \sqrt{\frac{8{\lambda }_{\mathrm{PCM}}\Delta tT_m-T_w}{{\rho }_{\mathrm{PCM}}{\gamma }_{\mathrm{PCM}}}}& \left(13\right)\end{array}$

The Power of the blower motor generating the cooling airflow (e.g., wind) is given by:

$\begin{array}{cc}N=\frac{169.6S}{T}u^3& \left(6\right)\end{array}$

using equation (5), we have:

$\begin{array}{cc}N=0.019{\mathrm{ST}}^{-1}{d^{1.76}\left(\frac{\Delta t_1\Phi }{\Delta t_2\Delta T_mA}\right)}^{4.76}& \left(7\right)\end{array}$

where N is the power of the air blower in units of W; S is the cross sectional area of the air duct in units of m²; T is the temperature of the ambient air in units of K; A is the heat transfer area in units of m² which is the sum of all the surface areas for the thermal storage package pipe (container) outside surfaces.

In certain embodiments, as indicated above the devices described herein can be configured in a “parallel” architecture comprising multiple devices, e.g., as illustrated in FIG. 2. In certain embodiments N, the number of devices in such a configuration is greater than 2, more preferably greater than 3, still more preferably greater than 4, 5, or 6, in certain embodiments, greater than 10, 20, 30, 40, or 50, and in certain embodiments, greater than 75, 100, 150 or greater than 200.

In this way, the system can handle larger amounts of exhausted steam (for example, larger than 3.6 MWh heat capacity, as in the example storage unit described above) or sustain 24 hours of continuing operation by alternating “condensing/storage” and “heat dissipation” process with different containers. Another advantage with this configuration is that the number of valves that allow ambient air to enter the tank during the dissipation cycle can be reduced significantly to lower the system cost. For example, in certain embodiments, the N containers can share two butterfly valves, one for air entering path and the other for the air exhaust path through the main pipe and parallel pipes configured as described in FIG. 2. In addition, one wind blower with larger operation power can be used for the entire system, further simplifying system construction and reducing cost.

Continuing with the example described above, and considering the heat dissipation stage. Table 1 lists typical heat dissipation time periods required for each thermal storage tank given the ambient air temperature for a typical year for the said 10 MW CSP power plant. The average ambient air temperatures for each month are for the application location in northern China. It should be noted, however, that because the ambient air temperature during the night is much lower than the phase change temperature relative to the temperature difference of T_w and T_m, the required Δt₂ is actually much shorter than Δt₁. In other words, the “delay” is much more effective than the “prolonged”, especially during the winter season.

Table 1. Illustrates typical heat dissipation time periods given the ambient temperature for a typical year and realistic application of the system.

Phase	Phase		Heat of	Thermal
change	change	density	fusion	conductivity
material	temperature	(kg/m3)	(kJ/kg)	(W/m · K)
Na2C03 · 10H20	32° C.	1349 ( l )	267	0.54
		1447 ( s )
				Dissipation
	Ambient			temperature
	temperature	Δt2	Tw − Tm	difference
January	−16.1/−7.2° C.	2.37 h	2.6° C.	26.8° C.
February	−14.3/−3.4° C.	2.45 h		25.7° C.
March	−1.9/9.9° C.	3.18 h		18.6° C.
April	5.1/17.7° C.	3.81 h		14.6° C.
May	9.2/23.1° C.	4.37 h		12.1° C.
June	14.7/26.3° C.	5.08 h		9.1° C.
July	17.7/29.6° C.	5.58 h		7.4° C.
August	15.0/25.2° C.	5.09 h		9.1° C.
September	10.5/21.9° C.	4.44 h		11.6° C.
October	2.4/15.1° C.	3.54 h		16.1° C.
November	−4.6/5.3° C.	2.97 h		20.3° C.
December	−13.0/−3.4° C.	2.51 h		25.0° C.

It is very easy to control the system operation. During the “condensation stage”, because the heat exchange area is sufficiently large and the required heat exchange temperature difference is relatively small, as described in the above examples, no specific control is needed except to open the control valve 102 to start the “condensation stage” and to close this valve when the condensation stage” is finished. During this stage, the exhaust steam will first condensed onto the packaged thermal storage pipe surface while transferring the latent heat into the phase changing material with very small temperature difference. As time passes, this heat transfer temperature difference gradually increases. However, even at the end of the stage, this temperature difference is still relatively small. In the above described example, this value is only 2.6° C. For the “heat dissipation stage”, as described above, valve 102 is closed, and valve 114 is opened to let air into the storage tank until the pressure inside the tank is at or close to outside pressure. The two valves 106 in FIG. 1 or valves 205 and 206 in FIG. 2 are opened. The blower tis turned on, or where a cooling air stream is provided by ducted ambient wind, the ducts are open. When used, the power setting of the blower depends on the ambient temperature, as illustrated in the above example and Table 1.

In various embodiments low temperature phase change materials are preferred as the thermal storage material for the delay and prolonged dry-cooling applications. Means of determining suitable parameters for the phase change materials are provided above. Illustrative low temperature phase change materials believed to be suitable for the heat storage material include, but are not limited to those shown in Table 2.

TABLE 2
Illustrative low temperature phase change materials suitable for exhaust
steam condensation/thermal storage applications.
		Heat of
Phase Change Thermal Storage Material	Melting point	Fusion
(molecular formula)	° C.	(KWh/ton)
LiClO3•3H2O	8.1	70.3
ZnCl2•3H2O	10	/
K2HPO4•6H2O	13/14	30.3
45% CaCl2•6H2O + 55% CaBr2•6H2O	14.7	38.9
NaOH 3½ H2O	15/15.4	/
51-55% Cu(NO3)2•6H2O + 45-49%	16.5	69.4
LiNO3•3H2O
45-52% LiNO3•3H2O + 48-55%	17.2	61.1
Zn(NO3)2•6H2O
Na2CrO4•10H2O	18	/
KF•4H2O	18.5-19.0	64.2
FeBr3•6H2O	21/27	29.2
55-65% LiNO3•3H2O + 35-45%	24.2	63.9
Ni(NO3)2•6H2O
66.6% CaCl2•6H2O + 33.4%	25	35.3
MgCl2•6H2O
45% Ca(NO3)2•6H2O + 55% Zn	25	36.1
(NO3)2•6H2O
50% CaCl2•6H2O + 50%	25	26.4
MgCl2•6H2O
Mn(NO3)2•6H2O	25.8/25.5	38.9
48% CaCl2 + 4.3% NaCl + 0.4%	26.8	52.2
KCl + 47.3% H2O
50% CH3CONH2 + 50% NH2CONH2	27	45.3
CaCl2•6H2O	29	53.0
CaCl2•6H2O	29/29.2/29.6/	50.0
	29.7/30/29.9
LiNO3•3H2O	30	82.2
LiNO3•3H2O	30	52.5
67% Ca(NO3)2•4H2O + 33%	30	37.8
Mg(NO3)2•6H2O
47% Ca(NO3)2•4H2O + 53%	30	37.8
Mg(NO3)2•6H2O
60% CH3COONa•3H2O + 40%	31.5	62.8
CO(NH2)2
Na2CO3•10H2O	32	74.2
Na2SO4•10H2O	32.4	66.9
KFe(SO4)2•12H2O	33	48.1
Na2CO3•10H2O	33/32	74.2
LiBr2•2H2O	34	34.4
CaBr2•6H2O	34	33.3
Na2HPO4•12H2O	35.5/36/35/40	77.8
0.925 mol Ca(NO3)2•4H2O + 0.075	35.6	43.1
mol CaCl2•6H2O
Zn(NO3)2•6H2O	36/36.4/36.1	40.8
FeCl3•6H2O	37	61.9
Mn(NO3)2•4H2O	37.1	31.9
Na2HPO4•12H2O	40	77.5
50% CH3COONa•3H2O + 50%	40.5	70.8
HCONH2
CoSO4•7H2O	40.7	47.2
KF•2H2O	41.4/42	45.0
CH3COOK•1½H2O	42	/
MgI2•8H2O	42	36.9
CaI2•6H2O	42	45.0
K2HPO4•7H2O	45	40.3
K3PO4•7H2O	45	/
Zn(NO3)2•4H2O	45/45.5	30.6
Ca(NO3)2•4H2O	42.7/47	42.5
Mg(NO3)2•4H2O	47	39.4
Fe(NO3)3•9H2O	47	43.1
Na2HPO4•7H2O	48	/
Na2SiO3•4H2O	48	46.7
K2HPO4•3H2O	48	27.5
MgSO4•7H2O	48.5	56.1
Na2S2O3•5H2O	48/48-49/48.5	58.1
Ca(NO3)2•3H2O	51	28.9
61.5% Mg(NO3)2•6H2O + 38.5%	52	36.5
NH4NO3

The foregoing embodiments are intended to bee illustrative an not limiting. Using teachings provided herein, other variations will be available to those of skill in the art. For example, while the system in FIG. 1 is shown with a fan 107 to drive ambient air through the thermal storage system during the heat dissipation period in various embodiments, a fan is not required. For example, where the system is sited in locations having significant winds, particular at night, the system can comprise ducts to channel ambient winds through the thermal storage system thereby reducing or avoiding the power usage required by a fan.

Similarly, a valve 104 is shown to control application of vacuum to chamber 108, while a separate valve 114 is shown to introduce atmospheric pressure air into chamber 108. However, in certain embodiments, these need not be separate valves. Thus, for example valve 104 can be configured to switch between a vacuum source and atmospheric pressure air thereby obviating valve 114.

FIG. 1 also shows that in certain embodiments, the device can comprise an agitator/mixer configured to rotate perforated plates to which the chambers containing the thermal storage material are attached. However, numerous other configures to achieve such mixing will be recognized and available to one of skill in the art.

Thus, it is understood that the examples and embodiments described herein are for illustrative purposes only and that various modifications or changes in light thereof will be suggested to persons skilled in the art and are to be included within the spirit and purview of this application and scope of the appended claims. All publications, patents, and patent applications cited herein are hereby incorporated by reference in their entirety for all purposes.

Claims

1. A device for steam condensation and delayed dissipation of the heat produced by said condensation, said device comprising:

a combined condensation/thermal storage chamber, said chamber comprising:

one or more valved ports for receiving steam from a steam source;

one or more containers containing a thermal storage material;

one or more valved ports for applying a vacuum to said thermal storage chamber;

one or more valved ports for introducing ambient pressure air into said chamber;

one or more valved ports for removing condensed water from said chamber; and

a valve system operably coupling said chamber to a source of ambient temperature air.

2. The device of claim 1, wherein said one or more containers containing a thermal storage material is a plurality of containers each containing a thermal storage material.

3. The device of claim 1, wherein said thermal storage material is a phase change thermal storage material (PCM).

4. The device of claim 3, wherein said thermal storage material is a liquid/solid phase change thermal storage material.

5. The device of claim 3, wherein said thermal storage material comprises a material selected from the thermal storage materials shown in Table 2.

6. The device of claim 3, wherein said thermal storage material comprises Na2CO3.10H2O.

7. The device of claim 3, wherein said PCM contains glass microfibers or nanofibers.

8. The device of claim 1, wherein said device further comprises an apparatus to cause mixing of liquid phase thermal storage material in the containers containing the thermal storage material.

9. The device of claim 8, wherein said containers are attached to a structure frame and said apparatus comprises a motor configured to cause rotation of said structure frame and the attached containers.

10. The device of claim 1, wherein said one or more valved ports for applying a vacuum to said thermal storage chamber are operably coupled to a vacuum pump.

11. The device of claim 1, wherein said one or more valved ports for applying a vacuum to said thermal storage chamber and said one or more valved ports for introducing ambient pressure air into said chamber are controlled by separate valves.

12. The device of claim 1, wherein said one or more valved ports for applying a vacuum to said thermal storage chamber and said one or more valved ports for introducing ambient pressure air into said chamber are controlled by the same valve(s).

13. The device of claim 1, wherein said valve system operably coupling said chamber to a source of ambient temperature air comprises a butterfly valve.

14. The device of claim 1, wherein said source of ambient air is a fan and/or blower.

15. The device of claim 1, wherein said source of ambient air comprises one or more ducts configured to receive ambient wind.

16. The device of claim 1, wherein said device is one of a plurality of said devices configured in a parallel configuration.

17. The device of claim 1, wherein said one or more valved ports for receiving steam from a steam source are operably coupled to the low temperature steam output from a turbine.

18-19. (canceled)

20. A system for delayed heat dissipation from the condensation of waste steam, said system comprising a plurality of devices according to claim 1, configured in a parallel configuration.

21. The system of claim 20, wherein said system comprises at least 10, or at least 20, or at least 30, or at least 40, or at least 50, or at least 60, or at least 70, or at least 80, or at least 90, or at least 100, or at least 150, or at least 200 of said devices.

22. The system of claim 20, wherein said system is operably coupled to the low temperature steam output from a turbine.

23-24. (canceled)

25. A dry-cooling condensation method, said method comprising

receiving steam from a source of steam;

condensing the steam into water while transferring the latent heat of said steam into the latent heat of a thermal storage material; and

dissipating the latent heat from said thermal storage material at a later time when the ambient temperature is lower than the ambient temperature at the time the steam was condensed into water.

26. The method of claim 25, wherein said thermal storage material is a phase change thermal storage material (PCM).

27. The method of claim 25, wherein said thermal storage material is a liquid/solid phase change thermal storage material.

28. The method of claim 25, wherein said thermal storage material comprises a material selected from the thermal storage materials shown in Table 2.

29. The method of claim 25, wherein said thermal storage material comprises Na2CO3.10H2O.

30. The method of claim 25, wherein said source of steam is the steam output from a turbine.

31. The method of claim 30, wherein said turbine is in a power plant selected from the group consisting of a coal-fired power plant, a gas-fired power plant, a nuclear power plant, and a solar thermal power plant.

32. The method of claim 25, wherein said receiving and condensing comprises receiving and condensing during daylight hours.

33. The method of claim 32, wherein said receiving and condensing comprises receiving and condensing between noon and 3:00 pm.

34. The method of claim 25, wherein said dissipating comprise dissipating the latent heat from said thermal storage material during the late afternoon, and/or evening, and/or night.

35. The method of claim 25, wherein said method is performed using a device comprising:

a combined condensation/thermal storage chamber, said chamber comprising:

one or more valved ports for receiving steam from a steam source;

one or more containers containing a thermal storage material;

one or more valved ports for applying a vacuum to said thermal storage chamber;

one or more valved ports for introducing ambient pressure air into said chamber;

one or more valved ports for removing condensed water from said chamber; and

a valve system operably coupling said chamber to a source of ambient temperature air.

36. The method of claim 35, wherein said receiving and condensing comprises:

opening said one or more valved ports for applying a vacuum to said thermal storage chamber to reduce the ambient pressure in said thermal storage chamber;

opening said one or more valved ports for receiving steam to introduce steam from said steam source into said chamber, whereby said steam condenses transferring latent heat of steam into said thermal storage material; and

operating said one or more valved ports for removing condensed water from said chamber to return the condensed water to the system providing said steam source.

37. The method of claim 35, wherein said dissipating comprises:

restoring the pressure in said thermal storage chamber to atmospheric pressure;

operating said valve system operably coupling said chamber to a source of ambient temperature air to pass ambient temperature air through said thermal storage chamber to transfer heat from said thermal storage material to said air.

38. The method of claim 37, wherein passing ambient temperature air comprising operating a fan/blower to force air through said chamber.

39. The method of claim 37, wherein passing ambient temperature air comprising coupling said chamber to a duct system that channels wind through said chamber.

40. The method of claim 37, wherein said dissipating further comprises operating an apparatus to provide mixing of fluid thermal storage material in the chambers containing said fluid thermal storage material.

41. The method of claim 40, wherein said method comprises operating a motor to rotate a structure frame to which said chambers containing said thermal storage material are attached.

42. The method of claim 35, wherein said method is performed using a system comprising a plurality of said devices.

43. The method of claim 42, wherein said devices are configured in a parallel configuration.

44. The method of claim 42, wherein substantially all of the devices in said system perform said receiving and condensing at the same time.

45. The method of claim 42, wherein substantially all of the devices in said system perform said dissipating at the same time.

46. The method of claim 42, wherein a plurality of the devices perform said receiving and condensing at the same time others of the devices are dissipating.