US20040211184A1

US20040211184A1 - Convection towers for air cooled heat exchangers

Info

Publication number: US20040211184A1
Application number: US10/407,415
Authority: US
Inventors: Desikan Bharathan; Keith Gawlik
Original assignee: Midwest Research Institute
Current assignee: Midwest Research Institute
Priority date: 2003-04-04
Filing date: 2003-04-04
Publication date: 2004-10-28

Abstract

An exemplary heat exchange system includes a first stage including a first stage heat exchanger and a tower wherein the first stage heat exchanger exchanges heat between a fluid and air to thereby heat the air and generate air convection in the tower; and a second stage including a second stage heat exchanger and a powerable convection unit wherein the second stage heat exchanger receives the fluid from the first stage and exchanges heat between the fluid and air. An exemplary method includes transferring heat energy from a fluid to air using a first heat exchanger to cause air convection in a tower; and transferring heat energy from the fluid to air using a second heat exchanger and a fan. Various other exemplary systems and methods are also disclosed.

Description

CONTRACTUAL ORIGIN OF THE INVENTION

[0001] The United States Government has rights I this invention pursuant to Contract No. DE-AC36-99GO10337 between the U.S. Department of Energy and the National Renewable Energy Laboratory, a division of Midwest Research Institute.

FIELD OF INVENTION

The subject matter disclosed herein generally relates to air cooled heat exchangers and/or air cooled heat exchange systems.

BACKGROUND

Hot air rises naturally because of its lower density in a cool atmosphere (e.g., consider a hot air balloon); hence, a substantial amount of air can be induced to flow naturally upward in a properly designed flow passage. Natural draft-cooling towers operate using this principle at many power plants.

Further, a typical residence may use a chimney to cause flue gases (from furnaces and fireplaces) to escape into the atmosphere. The key difference between the residential and power plant applications is that the chimney for the residence is a lot shorter and is made to fit naturally within the height of the building.

Because the temperature of the exhaust air from a furnace is likely to be high, the shorter height of the chimney for a residence is generally adequate. However, for a power plant, the natural draft tower of a large diameter must typically rise to a height of about 120 m (e.g., about 400 ft).

A natural draft tower uses the heat contained in the hot air to effectively induce airflow over a plant's heat rejecting surfaces. A natural draft tower can eliminate use of fans to induce airflow and an associated need for parasitic power. However, the natural draft tower, as already noted, typically needs to be quite tall. Many sites do not offer the flexibility to use such a tower on account of general height restrictions on structures and their imposing visual impact. Further, natural-draft towers are normally built using concrete (e.g., thin-shell concrete, etc.), which can be a major cost item for the power plant that increases substantially with respect to tower height.

Forced draft heat exchangers are commonly used to avoid problems related to natural draft towers. Forced draft heat exchange systems are typically used with outdoor mounted condensing coils for air conditioners and power plants where they are placed either vertically or horizontally above the ground with fans either blowing or inducing airflow over the coils. Forced draft fans are also used in wet cooling towers. While forced draft units can be made compact, they require substantial parasitic power to run the fans. Forced draft units also do not take advantage of the potentially usable energy in the exhausted hot air.

With respect to water cooling, water for heat rejection is generally on the decline, especially in newer construction. Reasons for the decline in water use for heat rejection include lack of water availability, very long water-rights acquisition periods, and adverse impact of visual plumes (clouds) that arise from wet cooling towers. Indeed, more and more power plants are required to use air as a heat rejection medium. However, specific heat for air is quite low in comparison to that of water. Air is also substantially less dense than water. Hence, a large volume of air must be induced for heat rejection from power plants. With air-cooling emerging as a new trend in power plant construction, effective methods, heat exchangers and/or systems to reduce power consumption for air-cooling are becoming important in power plant development and/or other applications.

Various exemplary heat exchange systems and exemplary methods presented herein aim to balance aspects of natural convection tower with aspects of forced draft units. Further, various exemplary systems and exemplary methods presented herein aim to reduce parasitic power requirements.

SUMMARY

BRIEF DESCRIPTION OF THE DRAWINGS

Features and advantages of the described implementations can be more readily understood by reference to the following description taken in conjunction with the accompanying drawings. [0011]
FIG. 1 shows an exemplary two stage heat exchange system; [0012]
FIG. 2 shows plots of temperature versus energy for a substantially single component working fluid and a multi-component working fluid; [0013]
FIG. 3 shows a T-x-y diagram for a multi-component mixture of ammonia and water; [0014]
FIG. 4 shows a plot of temperature versus energy for an exemplary two stage heat exchange system that uses a working fluid having a temperature glide for a vapor to liquid phase transition; [0015]
FIG. 5 shows an exemplary heat exchange stage that includes a powerable convection unit; [0016]
FIG. 6 shows an exemplary heat exchange stage that includes a tower and optionally a powerable convection unit; [0017]
FIG. 7 shows an exemplary tower suitable for use in a heat exchange stage; [0018]
FIG. 8 shows an exemplary tower suitable for use in a heat exchange stage; and [0019]
FIG. 9 shows an exemplary two stage heat exchange system that relies on a structural passageway of a structure or building to help generate natural convection currents.[0020]

DETAILED DESCRIPTION

The following description includes the best mode presently contemplated for practicing various described implementations. This description is not to be taken in a limiting sense, but rather is made merely for the purpose of describing the general principles of the various implementations. The scope of the described implementations should be ascertained with reference to the issued claims. [0021]
Exemplary Heat Exchange System [0022]
FIG. 1 shows an exemplary two stage [0023] heat exchange system 100. A first stage, Stage 1, includes a heat exchanger 110 to exchange heat between a working fluid and air. In Stage 1, a tower 112 enhances air convection to thereby increase heat transfer between the working fluid and air at the heat exchanger 110.
According to [0024] Stage 1, a working fluid enters the heat exchanger 110 at a temperature T_WFS1(In) and air enters the heat exchanger 110 at a temperature T_AirS1(In) wherein T_WFS1(In) is greater than T_AirS1(In). At the heat exchanger 110, heat energy flows from the working fluid to the air. Depending on the state of the working fluid, the loss of heat energy normally causes the working fluid to experience at least a partial change of state (e.g., a phase change) and/or to decrease in temperature. For example, working fluid exiting the heat exchanger 110 may have a temperature T_WFS1(Out), which is less than T_WFS1(In).
Heat energy transferred to the air normally causes the air to increase in temperature. Hence, air exiting the [0025] heat exchanger 110 will have a temperature T_AirS1(Out), which is greater than T_AirS1(ln). In instances where the air entering the heat exchanger is taken from an ambient source, such as, air near the tower 112, heat energy transferred to the air at the heat exchanger 110 causes the air to expand and become less dense than the ambient air. In turn, the heated air rises in the tower 112 according to principles of natural convection. Thus, natural convection currents enhance heat transfer at the heat exchanger 110.
A second stage of the two stage [0026] heat exchange system 100, Stage 2, includes a heat exchanger 120 to exchange heat between the working fluid and air. In Stage 2, an air duct 122 provides a flow path for air past or through the heat exchanger 120. In this example, a fan 124 consumes power 126 to propel air or cause air convection in the duct 122. Air convection caused by power 126 provided to the fan 124 aims to increase heat transfer between the working fluid and air at the heat exchanger 120.
According to [0027] Stage 2, the working fluid enters the heat exchanger 120 at a temperature T_WFS2(In) and air enters the heat exchanger 120 at a temperature T_AirS2(In) wherein T_WFS2(In) is greater than T_AirS2(In). At the heat exchanger 120, heat energy flows from the working fluid to the air. Depending on the state of the working fluid, the loss of heat energy normally causes the working fluid to experience at least a partial change of state (e.g., a phase change) and/or to decrease in temperature. For example, working fluid exiting the heat exchanger 120 may have a temperature T_WFS2(Out), which is less than T_WFS2(In).
Heat energy transferred to the air normally causes the air to increase in temperature. Hence, air exiting the [0028] heat exchanger 120 will have a temperature T_AirS2(Out), which is greater than T_AirS2(In). In instances where the air entering the heat exchanger is taken from an ambient source, heat energy transferred to the air at the heat exchanger 120 causes the air to expand and become less dense than the ambient air. In turn, the heated air may rise according to principles of natural convection. Thus, natural convection currents may enhance heat transfer at the heat exchanger 120. However, in this example, Stage 2 consumes power from the power source 126 to enhance heat exchange at the heat exchanger 120. In general, such power may be referred to as parasitic power. Parasitic power demands may depend on properties of the working fluid and/or working fluid flow requirements. For example, a requirement may be placed on T_WFS2(Out), which corresponds to a certain vapor/liquid phase composition of the working fluid at a given working fluid pressure.
While the exemplary two stage [0029] heat exchange system 100 shows a single heat exchanger 110 and a single tower 112 in Stage 1, one or more heat exchangers and/or one or more towers may be used in alternative examples. Further, while the exemplary two stage heat exchange system 100 shows a single heat exchanger 120, a single duct 122, a single fan 124 and a single power source 126 in Stage 2, one or more heat exchanges, one or more ducts, one or more fans, one or more power sources, and/or one or more towers may be used in alternative examples. In general, Stage 1 relies predominantly on natural convection and Stage 2 relies, at least in part, on a power source.
FIG. 2 shows a set of temperature versus energy plots [0030] 200. A first temperature versus heat load plot 210 corresponds to a substantially single component working fluid whereas a second temperature versus heat load plot 220 corresponds to a multi-component working fluid. In these plots, heat load is given in energy per unit time.
In the [0031] plot 210, the segment A-B corresponds to heating a first phase of the substantially single component working fluid from a temperature to a phase transition temperature. The energy input for sensible heating over the segment A-B is approximated by the equation Q_S-Liq.=m_Liq.C_Liq.ΔT_Liq., wherein m_Liq.is the mass of the working fluid, C_Liq.is the specific heat capacity for the first phase, and ΔT_Liq.is the temperature differential over the segment A-B.
In the [0032] plot 210, the segment B-C corresponds to heating a first phase of the substantially single component working fluid to effectuate a phase transition to a second phase of the substantially single component working fluid. In this example, the phase transition occurs at an approximately constant temperature. The energy input for latent heating over the segment B-C is approximated by the equation Q_L=m_VL_V, wherein m_Vis the mass of working fluid and L_Vis the associated latent heat energy per unit mass for the phase transition.
In the [0033] plot 210, the segment C-D corresponds to heating a second phase of the substantially single component working fluid from a phase transition temperature to a higher temperature. The energy input for sensible heating over the segment C-D is approximated by the equation Q_S-V=m_VC_VΔT_V, wherein m_Vis the mass of the working fluid, C_Vis the specific heat capacity for the first phase, and ΔT_Vis the temperature differential over the segment C-D. The total energy transferred to the working fluid over the segment A-D is approximated by the sum of energies associated with the individual segments: Q_T=Q_L+Q_S-V+Q_S-Liq.
As already mentioned, the [0034] plot 220 corresponds to a multi-component working fluid. In general, a suitable multi-component working fluid includes components that are miscible. For example, oil and water are typically not miscible under standard conditions (e.g., standard temperature, pressure, gravity, etc.). Further, a suitable multi-component working fluid may be a zeotrope or an azeotrope.
For a zeotrope, individual component concentrations in a liquid phase and a corresponding vapor phase are never equal, which creates a temperature glide during a liquid-vapor or vapor-liquid phase transition. In essence, a zeotrope is a liquid mixture that exhibits no maximum or minimum when vapor pressure is plotted against composition at constant temperature. As a zeotrope vapor is cooled, liquid formation commences at the dew point temperature and stops at the bubble point temperature. The difference between the dew point and bubble point temperatures is known as a temperature glide. In contrast, the substantially single component working fluid shown in [0035] plot 210 does not exhibit a temperature glide over the phase transition represented by the segment B-C because the phase transition occurs at an approximately constant temperature.
An azeotrope may be defined as a multi-component mixture at a point wherein the individual component concentrations of the liquid phase and the vapor phase are the same for a given temperature and pressure. An azeotrope typically behaves like a single component fluid in that a phase transition occurs at an approximately constant temperature. For example, an ethanol and water mixture is an azeotrope at approximately 0.89 mole fraction of ethanol, a pressure of approximately 101 kPa and a temperature of approximately 78.1° C. Thus, distillation of such a mixture can never produce a pure ethanol vapor. [0036]
Referring again to FIG. 2, [0037] plot 220 shows a temperature glide ΔT_Glideover a segment B-C. In addition, the plot 220 shows segments A-B and C-D, which correspond to liquid and vapor heating, respectively. The energy associated with heating the multi-component working fluid from A-D may still be represented as the sum of individual sensible and latent energies: Q_T=Q_L+Q_S-V+Q_S-Liq.
FIG. 3 shows a [0038] plot 300 of temperature versus mass fraction for an ammonia-water zeotrope at approximately 4 bar and at approximately 2 bar. Two pressures are shown to illustrate the effect of increasing pressure on vapor and liquid phase composition in terms of mass fraction. An exemplary condensation scheme is shown for an ammonia-water zeotrope having a liquid composition of approximately 0.35 mass fraction ammonia and approximately 0.65 mass fraction water. Of course, other compositions may be suitable. For example, an exemplary alternative ammonia-water zeotrope has a liquid composition of approximately 0.85 mass fraction ammonia and 0.15 mass fraction water. Yet further, other pressures may be possible, for example, an exemplary alternative uses an ammonia-water zeotrope at a pressure of approximately 13 bar (e.g., approximately 1.3 MPa or approximately 190 psia).
According to the exemplary condensation scheme, zeotrope vapor at a pressure of approximately 2 bar and a temperature of approximately 390 K (approx. 117° C.) (e.g., point D) losses energy to its surroundings. Upon reaching its dew point (e.g., point C), condensation commences wherein vapor and liquid or condensate concentrations differ. As more energy is lost, condensate concentration follows a line of bubble points while vapor concentration follows a line of dew points. At the zeotrope's bubble point (e.g., point B), the zeotrope becomes a single phase liquid. During condensation, the segment C-B, zeotrope temperature decreased according to a temperature glide ΔT[0039] _Glide. A further loss of energy causes the zeotrope to cool as a liquid to a temperature of approximately 310 K (approx. 37° C.). Thus, in this example, the overall temperature differential, ΔT, is approximately 80 K. Of course, other temperature differentials may be suitable and depend on zeotrope concentrations, pressure, etc. For example, an exemplary zeotrope operates with a temperature differential of approximately 30 K and between approximately 343 K (approx. 70° C.) and approximately 313 K (approx. 40° C.).
Zeotropes often exhibit a nonlinear relationship between enthalpy and temperature whereas a single component typically exhibits a linear relationship. Consequently, a zeotrope may have a specific heat capacity that is nonlinear with respect to temperature. For example, an ammonia-water zeotrope may have a specific heat capacity that increases nonlinear at temperatures greater than approximately 100° C. and that becomes more nonlinear with respect to an increase in the mass fraction of ammonia. [0040]
FIG. 4 shows an [0041] exemplary plot 400 of temperature versus heat load for a working fluid (dashed line) and air (solid lines). In the plot 400, heat load is given as energy per unit time. The information in the plot 400 may be related to the exemplary two stage heat exchange system 100 of FIG. 1, the exemplary working fluid behavior shown in plot 220 of FIG. 2 and/or the exemplary working fluid behavior shown in plot 300 of FIG. 3.
In [0042] Stage 1, a working fluid at a temperature T_WFS1(In) (e.g., point D) exchanges heat with air at a temperature T_AirS1(Out). The working fluid cools as a vapor until it reaches point C, which corresponds to a dew point of the working fluid. Thus, in this example, at least some condensation occurs during Stage 1. The working fluid continues to transfer heat energy to the air until point S_1-2, which coincides with air at a temperature T_AirS1(In) and working fluid at a temperature T_WFS1(Out). While the plot 400 shows air heating and working fluid cooling in a substantially countercurrent manner, actual heat transfer may occur in other manners depending on heat exchanger arrangement and/or the number of heat exchangers in Stage 1. The plot 400, for Stage 1, primarily indicates that a working fluid heats air to increase air temperature by a temperature differential ΔT_AirS1.
Working [0043] fluid exiting Stage 1 has a temperature T_WFS1(Out) (e.g., at point S_1-2). In general, the working fluid entering Stage 2 has a temperature T_WFS2(In), which is approximately equal to T_WFS1(Out). In Stage 2, the working fluid transfers heat energy to air to increase air temperature by a temperature differential ΔT_AirS2, which is the difference between T_AirS2(Out) and T_AirS2(In). During Stage 2, the working fluid continues to condense until reaching point B, which corresponds to a bubble point of the working fluid. Thereafter, cooling of the working fluid continues until reaching a temperature T_WFS2(Out) (e.g., point A). In Stage 2, at least some parasitic power is used to enhance heat transfer between the working fluid and the air. For example, as shown in FIG. 1, a power source 126 supplies power to a fan 124 that drives air in a duct 122 whereby the air passes near or through a heat exchanger 120.
FIG. 5 shows an axisymmetric cross-sectional view of an [0044] exemplary Stage 2 heat exchange system 500 wherein a central axis coincides with a z-dimension. The system 500 includes a duct 510, a heat exchanger 520, a powerable convection unit 530 and a power source 535 that can supply power to the convection unit 530. The duct 510 has an air inlet 512 at a position r_iand having a height Δz_iand an air outlet 514 at a position z_oand having a radius r_o. The heat exchanger 520 receives working fluid at a temperature T_WFS2(In) and cools the working fluid to an exit temperature T_WFS2(Out) using air provided at a temperature of T_AirS2(In). Heated air exits the heat exchanger 520 at a temperature T_AirS2(Out). The heated air flows in the duct 510, through the powerable convection unit 530 and exits the duct air outlet 514.
A trial was conducted using a system that simulated and approximated a heat exchange system such as the [0045] exemplary Stage 2 heat exchange system 500. To simulate a heat exchanger, a circular unit heater (Young Radiator Inc., manufacturer, Model AV-60S) was used. The heater consisted of a set of three parallel 16-mm (0.625-inch) copper coils fitted with aluminum fins. The coils were wound into a circular form. Inlet and outlet manifolds for the coils were nominal 25-mm (1-inch) steel pipes. Air was drawn circumferentially through the coil by an axial fan located centrally. Portions were masked circumferentially to assure air flowed through the fins and there was no other leakage or bypass paths for the airflow. The exhaust nozzle was 356 mm (14 inch) in diameter.
FIG. 6 shows an axisymmetric cross-sectional view of an [0046] exemplary Stage 1 heat exchange system 600 wherein a central axis coincides with a z-dimension. The system 600 includes a duct 610, a heat exchanger 620, an optional powerable convection unit 630 and an optional power source 635 that can supply power to the optional convection unit 630. The duct 610 has an air inlet 612 at a position r_iand having a height Δz_iand an air outlet 614 at a position z₀and having a radius r_o. The heat exchanger 620 receives working fluid at a temperature T_WFS1(In) and cools the working fluid to an exit temperature T_WFS1(Out) using air provided at a temperature of T_AirS1(In). Heated air exits the heat exchanger 620 at a temperature T_AirS1(Out). The heated air flows in the duct 610, through the powerable convection unit 630, if provided, and exits the duct the air outlet 614. In this example, the duct air outlet 614 connects to a tower 640. The tower 640 receives heated air at the duct outlet 614 and expels the heated air at a tower air outlet 644. The tower 640 has a tower height Δz_Toand a tower air outlet radius of r_To. The tower air outlet radius r_Tomay differ from the duct outlet radius r_o.
A trial was conducted wherein the aforementioned system that simulated and approximated a heat exchange system was fitted with an insulated flexible duct having a diameter of approximately 356 mm (14 inch diameter), which received air from the exhaust nozzle and simulated a tower. [0047]
The simulated tower duct was hung vertically from the ceiling of a building. Trials were conducted using various heights for this simulated tower (e.g., Δz[0048] _TO).
In the trials, steam at atmospheric pressure was used to provide heat to the coils. The heater was tested with and without use of the fan to induce airflow. For tests without the fan, the fan blades were removed from the motor shaft. The electric motor powering the fan was rated at nominally 100 W and rotated at speed of 1090 rpm. [0049]
Instrumentation for the trials included type-K thermocouples for the air temperature measurements and an inclined tube manometer for the measuring airside pressure drops. Barometric pressures were also noted during the trials. The rejected heat rate was inferred by collecting and weighing steam condensate over measured elapsed period of time using a stopwatch. [0050]
During trials, air entered the system radially near the bottom and then traveled upward through a plastic transition section to the simulated tower. The simulated tower included outer insulation having Mylar reflective film on outer surface of the insulation. [0051]
Infrared images were taken of the system and an exiting air plume. To visualize better the plume, a sheet of drawing paper was placed in the path of the exiting air plume. A colored temperature field was indicative of the velocity profile of the exiting air. During operation, the heater was the hottest and brightest area in the infrared images and was held at a temperature of approximately 95° C. [0052]
Measured temperature rise in the air together with the rejected heat rate were used in calculations to derive a value for airflow. These data were then converted to variations of pressure loss and heat transfer coefficients as functions of airflow. [0053]
Measured pressure loss (e.g., in Pa) across the coil as a function of the incoming air volumetric flow rate (e.g., in m[0054] ³/s) was plotted. The plot showed that pressure loss increased substantially monotonically with increasing flow. A comparison to calculated variations in pressure loss with respect to flow rate, based on laminar flow between parallel fins, showed that the trial measurements were quite close to the calculated values.
Measured overall heat transfer coefficient (e.g., in W/m[0055] ²K) was plotted as a function of incoming air volumetric flow rate (e.g., in m³/s). The coefficient was based on the actual heat transfer area available, rather than the tube external surface area. The measured values fell between approximately 60 W/m²K and approximately 90 W/m²K. Accordingly, the heat transfer coefficient increases with increasing flow rate. At low flow rates, a laminar flow limit (for constant temperature boundary condition) occurred near 50 W/m²K for flow rates to approximately 0.8 m³/s (corresponding to a Reynolds number of 2000 based on fin spacing). A corresponding turbulent limit was determined at the higher flow rates, based on Dittus-Boelter empirical correlation. Both predictions were for fully developed thermal boundary layers.
The trials showed that for the tested fins with a length-to-hydraulic-diameter ratio of approximately 11, flow does not develop fully. Consequently, under such circumstances, the measured heat transfer coefficients tend to be larger than those predicted, as borne out by the trial data. The trials helped to establish a set of base line data suitable for use in determining flow and heat transfer performance for air-cooled heat exchangers. [0056]
Further trials used a computational fluid dynamics (CFD) model of an exemplary heat exchange system wherein tower heights were varied. For the CFD model, commercially available software was used (FLUENT®, Aavid Thermal Technologies, Inc., New Hampshire). All models used axisymmetric flow to reduce model complexity. Flow geometries and meshes were generated using accompanying software. Proper boundaries including the symmetry axis, walls, the heater and fan areas were identified at mesh generation. The mesh was then imported into the software program. Airflow was modeled using ideal gas theory. Buoyancy forces were accounted for by imposing gravitational forces along a downward direction parallel to the axis of any particular exemplary tower. [0057]
The CFD software yielded velocity profiles that were induced in the neighborhood of the heat exchanger. Three different velocity profiles corresponding to three different tower section heights indicated that maximum velocity increases with increased tower section height. In particular, tower heights of approximately 2.5 m, 4.4 m, and 5.3 m were used wherein velocity magnitude varied from 0 m/s (e.g., at walls and the bottom) to approximately 10 m/s at the tower air outlet. [0058]
Exemplary Power Plant Cooling System [0059]
As described herein, a geothermal power plant is used to demonstrate various exemplary heat exchange systems and, in particular, an exemplary two stage heat exchange system such as the [0060] exemplary system 100 of FIG. 1. First, a geothermal power plant is chosen for evaluation. Next, the plant is evaluated for a single stage heat exchange system without a tower. Then, the plant is evaluated for an exemplary single stage heat exchange system with a tower. Finally, the plant is used to demonstrate an exemplary two stage heat exchange system.
A nominal 12.5 MW net geothermal power plant was evaluated for the use of an exemplary two stage heat exchange system. In this evaluation, the plant included a bottoming binary power system that was added to an existing flash power plant, such as at the Blundell geothermal power plant site in Utah. Many geothermal power plants, on account of the low temperature resource when compared to fossil plants, reject about 85% to 90% of incoming heat. Therefore, a substantial heat rejection system is required. Normally such a heat rejection system occupies a considerable portion of the plant site. [0061]
For this evaluation, the bottoming plant uses a non-azeotropic mixed working fluid (e.g., a zeotrope). For example, the bottoming plant may use a mixture of ammonia and water as a working fluid for the power system and may use an air-cooled heat rejection system. In general, ammonia-water working fluids are suitable for plants that rely on Kalina power cycles. When compared to a Carnot cycle (see, e.g., discussion of most single component working fluids), a non-azeotropic working fluid in a Kalina power cycle vaporizes and condenses at varying temperatures (e.g., according to a temperature glide) to result in increased work yield. Other cycles that may be used include Rankine cycles, Uehara cycles and Lorenz cycles. [0062]
According to the evaluation plant, an ammonia-water zeotrope (e.g., approximately 84.62% ammonia by mass) enters a heat exchanger at a flow rate of approximately 98 kg/s, at a temperature of approximately 66° C., a pressure of approximately 1.3 MPa and a vapor fraction of approximately 0.67. Ambient air enters the heat exchanger at a flow rate of approximately 5720 kg/s, a temperature of approximately 32° C. and a pressure of approximately 81 kPa. The air cools and condenses the working fluid to a temperature of approximately 40° C. In turn, the air heats to a temperature of approximately 47° C. In this example, approximately 86 MW of power is rejected (e.g., heat energy per unit time). [0063]
In this example, the one stage heat exchange system causes an airside pressure loss of approximately 140 Pa. For an assumed overall fan efficiency of 0.54, the required fan parasitic power for inducing the associated airflow is approximately 1.56 MW. This amount of power represents a loss of 12.5% of the net power on fans. [0064]
To reduce this parasitic power, an exemplary one stage heat exchange system with a tower is implemented. Based on modeling results, the evaluation plant requires a heat exchange system having an overall UA of approximately 11×10[0065] ⁶W/K. Assuming a nominal air-side heat transfer coefficient (based on bare tube area) of approximately 625 W/m²K, the tube area requirement is approximately 17,600 m². Based on data provided in various sources, a conventional air-cooled condenser with the fans has a cost of approximately $3.2 million (equipment cost—non-installed cost). This cost may represent nearly one third of the total equipment cost wherein the overall installed plant cost is estimated to be approximately $28 million.
For purposes of demonstrating an exemplary one stage heat exchange system having a tower, the evaluation plant is modeled after the Blundell Geothermal Power Plant located approximately 300 km (e.g., 250 miles) south of Salt Lake City, Utah. The plant has a dual flash power system with a nominal 23.5 MW gross output. On account of potential precipitation from the exit brine, the plant discharges brine at a temperature of approximately 170° C. The plant is situated on a gradually sloping ground with neighboring mountains approximately 1 km (e.g., 0.5 mile) to the east. The neighboring mountain range rises to a height of about 150 m (e.g., 500 ft) above the plant. Based on the terrain of the evaluation plant, a tower height of approximately 60 m was chosen as to be not too obtrusive and to still provide a reasonable height. [0066]
Based on the inlet and exit air temperatures to the heat exchanger, such a tower could provide a pressure difference of approximately 30 Pa (0.12 inch of water). The potential parasitic fan power reduction on account of the tower would be nominally one-fourth of its value without the tower. This savings correspond to 375 kW. [0067]
Of course, a taller tower and/or a hotter air exit temperature would increase the potential benefit of such an exemplary one stage heat exchange system having a tower, with the power reduction being directly proportional to the product of these variables. [0068]
FIG. 7 shows an [0069] exemplary tower 700 optionally suitable for use in the exemplary one stage heat exchange system and/or various other exemplary heat exchange systems described herein. To demonstrate use of the exemplary tower, a heat exchanger (or the condenser) was sized with a nominal inlet superficial air velocity of approximately 2 m/s (e.g., at a tower air inlet 712) and a tower exit velocity of approximately 5 m/s (e.g., at a tower air outlet 714). To handle 5610 m³/s of inlet air volumetric flow, the heat exchanger frontal area is assumed to be approximately 2805 m². In this example, the heat exchanger was arranged around the periphery of a substantially cylindrical air inlet section 712 of the exemplary tower 700. Given the requirements, the heat exchanger overall diameter and height were approximately 75 m and approximately 12 m, respectively. For this example, the height of the exemplary tower 700 was approximately 60 m and the exit diameter of air outlet 714 was approximately 40 m, which resulted in an exit air velocity of approximately 4.5 m/s.
The [0070] exemplary tower 700 optionally includes a flow guide 718, which has a varying radius that is larger near the bottom of the tower (e.g., near the air inlet 712) and smaller at axial positions the air inlet 712. The flow guide 718 may form an internal column along the axis of the tower 700. In one example, the flow guide 718 forms a column having a radius of approximately several meters. The exemplary tower optionally includes a powerable convection unit 730, for example, a unit that includes fans to help draw air into the air inlet 712.
In one example, the [0071] powerable convection unit 730 includes fans to induce the airflow positioned at a height of approximately 20 m (e.g., about one-third the tower height). To induce the required airflow, six fans were arranged peripherally around the axis of the tower. For example, fans such as those available from Hudson Air Products, Inc., having a diameter of approximately 12 m. To provide access to associated fan machinery, the exemplary tower 700 has a tower radius of approximately 18 m at the height of the convection unit 730.
Overall, in this particular example, the [0072] exemplary tower 700 has a total height of approximately 60 m and effective height of approximately 55 m, which can induce a buoyancy pressure differential of approximately 24 Pa. The pressure differential aids convection and hence reduces powerable convection unit requirements. In this example, the parasitic power loss for the powerable convection unit 730 (e.g., the fans) is approximately 1291 kW, which results in a power savings of 270 kW compared to a heat exchange system without a tower.
While the [0073] exemplary tower 700 has a substantially cylindrical tower section, polygonal tower sections are also possible (e.g., 6-sided, 12-sided, etc.) and may ease construction and reduce costs. Construction material is optionally concrete (e.g., thin-shell, etc.) and/or other material.
As already mentioned, the evaluation plant is suitable for demonstrating an exemplary two stage heat exchange system, such as, but not limited to, the exemplary [0074] heat exchange system 100 of FIG. 1. In essence, the two stage approach splits the cooling requirements into a first stage and a second stage. At times, the first stage is referred to as a topping condenser and the second stage is referred to as a bottoming condenser. The first stage aims to take advantage of the highest post-expansion temperature of the working fluid. For example, if the working fluid is exiting a turbine wherein it expands to drive the turbine, residual heat energy in the working fluid is at its highest value just after exiting the turbine. Of course, due to plumbing and/or other considerations, some heat energy may be lost prior to contacting air in a heat exchange system having a tower.
According to an exemplary method, airflow to a first stage of an exemplary heat exchange system is adjusted to allow the air temperature to reach a maximum given a pinch point constraint (e.g., accounting for second stage requirements and/or operational parameters). Of course, a natural equilibrium may be reached wherein the convection forces balance heat transfer to the air. For example, an increase in air temperature causes an increase in convection, which, in turn, may increase air flow past or through a heat exchanger. The increase in air flow past or through the heat exchanger may aid cooling of a working fluid; however, it will decrease residence time of the air and hence result in a lower air temperature. In turn, the lower air temperature will cause a decrease in natural convection. [0075]
Various exemplary methods described herein may be implemented using an exemplary controller or control system. In general, a controller allows for control of parameters germane to heat exchange (e.g., pressure drop, working fluid flow rate, working fluid pressure, air flow rate, air pressure, parasitic power where applicable, etc.). [0076]
An exemplary first stage includes a heat exchanger positioned around the periphery of an exemplary tower wherein air convection occurs due to natural convection. An alternative exemplary first stage includes a powerable convection unit, which may be used continuously and/or intermittently. Such a powerable convection unit is optionally used for control and/or for emergency cooling. Where an exemplary first stage does not include a powerable convection unit (e.g., a fan, etc.) or wherein no power flows to a powerable convection unit, there is no convection-related parasitic power loss associated with transfer of heat from a working fluid to air. [0077]
According to an exemplary two stage heat exchange system, working fluid exiting the first stage is received by the second stage (e.g., by a bottoming condenser, etc.). Of course, various exemplary schemes are possible whereby working fluid condensed in the first stage is diverted and only remaining vapor fed to the second stage. [0078]
In this particular example, the second stage uses a powerable convection unit (e.g., one or more fans, etc.) that consumes power (i.e., parasitic power) to force air past or through one or more heat exchangers. In this manner, the highest temperature working fluid is used in a stage to generate natural convection currents that enhance heat transfer while a lower temperature working fluid is cooled using at least some parasitic power. Again, in an exemplary multistage heat exchange system, constraints may exist as to pinch points or other parameters germane to operation of any individual stage or an entire power plant. [0079]
Referring again to the evaluation plant examples, an exemplary two stage heat exchange system has a first stage UA requirement of approximately 4.3 MW/K and a second stage UA requirement of approximately 7.5 MW/K. The (UA) requirement translates to an estimated pressure loss for each stage: approximately 52 Pa for the first stage and approximately 93 Pa for the second stage. In this example, the exemplary two stage heat exchange system results in a parasitic power reduction of approximately 1060 kW, which may be associated with operation of the second stage (e.g., a bottoming condenser, etc.). As such, the exemplary two stage heat exchange system results in a saving of approximately 500 kW. [0080]
Referring to FIG. 4, a similar plot exists for this exemplary two stage heat exchange system. For example, ambient air at a temperature of approximately 32° C. enters a stage one heat exchanger and is heated by a fluid to a temperature of approximately 60° C. The heated air rises in a tower, which, in turn, draws more ambient air into stage one. During this stage one air heating process, the fluid temperature decreases from approximately 67° C. to approximately 47° C. The associated heat duty decreases from approximately 87,000 kW to approximately 50,000 kW. For the second stage, ambient air enters at a temperature of approximately 32° C. and is heated by the fluid exiting stage one to a temperature of approximately 40° C. During this stage two air heating process, the fluid temperature decreases from approximately 47° C. to approximately 40° C. The associated heat duty decreases from approximately 50,000 kW to approximately 0 kW. Depending on pressure and/or composition of the fluid, vapor cooling, condensation and/or liquid cooling occurs during stage one and/or stage two. For example, stage one may include vapor cooling to a dew point followed by condensation along a temperature glide while stage two may include condensation along a temperature glide followed by liquid cooling. [0081]
Again, in this example, the total UA requirement for the two stages is approximately 11.7 MW/K, which is only slightly larger than the baseline requirement of 11 MW/K. The added area requirement in this example may be accounted for in an economic evaluation. [0082]
Overall, various aforementioned examples demonstrate utility of a multi-stage approach to cooling a working fluid. In particular, several examples demonstrate that high temperature, post-expansion working fluid is useful to generate natural convection to thereby reduce parasitic power requirements. Yet further, an exemplary two stage heat exchange system that includes powered convection in a second stage and natural convection in a first stage demonstrates reduced parasitic power requirements when compared to a single stage powered convection heat exchange system. [0083]
Exemplary Tower [0084]
FIG. 8 shows a rendered view cut open to show the internal features of an [0085] exemplary tower 800. The exemplary tower 800 has a sheath 810 that is supported internally and/or externally by a support system 816. The tower 800 includes an air inlet 812 and an air outlet 814. In a specific example, the exemplary tower 800 includes 12 vertical panels, 12 radial cables to hold the structure outward, connected by horizontal connector cables to ring cables that are spaced apart vertically. The panels (e.g., fabric or other material) may form a flow channel and have cables in cuffs along vertical edges. In this example, the edges are shaped to connect to the cable net at the intersection of ring cables and horizontal connector cables. The panels are optionally attached with industrial loop and hook attachments that can provide an airtight, smooth, continuously curved surface to enclose the air passage. The horizontal cables may permit accurate shaping of the fabric to follow the desired shape.
In this example, the edge supports consist of two sloping pipes that carry the load of the tower to the ground. The central mast includes a vertical section of space frame anchored at the bottom. Twelve stabilizing cables anchor the top ring to the mast foundation. The weight of the [0086] exemplary tower 800 is minimal compared to conventional cooling towers while the exemplary tower 800 has integrity sufficient to withstand cross winds.
Exemplary Multi-stage Heat Exchange for Buildings/Structures [0087]
Many tall buildings have the potential to make use of their natural height to help minimize the fan power used for heat rejection. Internal heat load from a building usually ends up as rejected condenser heat (along with the compressor work) from air conditioners. Ambient airflow is induced through the condenser coil using fans to carry the heat. In a typical application, hot air from a heat rejection system may exit with an air temperature rise of approximately 30° C. to approximately 45° C., and a corresponding decrease in density of approximately 10% to approximately 15%. Using an average value of 12% change in density, the buoyancy of hot air causes a potential motive pressure difference of about 1.44 Pascal per meter draft tower height. If an air cooler causes a pressure loss of approximately 125 Pa, the entire pressure loss can be overcome using a tower of approximately 87 m height. Thus, a tall building of 25 stories may induce all the airflow using natural convection. Of course, a shorter building will induce a lesser pressure difference for the airflow. [0088]
One suitable place to accommodate a draft tower is in a stairwell of a building. Stairwells are required for most all buildings to meet the safety related regulations. For example, an exemplary structural passageway is formed as a central clear column in a stairwell with stairs located around the periphery of the column. Such a column can be used as a draft tower to allow hot air to rise and thereby generate natural convection currents. Of course, hot air should be generated from heat exchange equipment located below, at or near the ground level of a building and directed to the bottom of such a structural passageway. [0089]
At early stages of planning, adequate provisions may be made to accommodate a structural passageway that acts as a draft tower and to provide for reasonable access to heat-rejection equipment for maintenance purposes. Whether the added costs related to such a structural passageway balances potential cost savings projected for fan use is likely to be application dependent. [0090]
FIG. 9 shows an exemplary two stage [0091] heat exchange system 900 in a structure 902 (e.g., a building, etc.). A first stage, Stage 1, includes a heat exchanger 910 to exchange heat between a working fluid and air. In Stage 1, a structural passageway 912 enhances air convection to thereby increase heat transfer between the working fluid and air at the heat exchanger 910.
According to [0092] Stage 1, a working fluid enters the heat exchanger 910 at a temperature T_WFS1(In) and air enters the heat exchanger 910 at a temperature T_AirS1(In) wherein T_WFS1(In) is greater than T_AirS1(In). At the heat exchanger 910, heat energy flows from the working fluid to the air. Depending on the state of the working fluid, the loss of heat energy normally causes the working fluid to experience at least a partial change of state (e.g., a phase change) and/or to decrease in temperature. For example, working fluid exiting the heat exchanger 910 may have a temperature T_WFS1(Out), which is less than T_WFS1(In).
Heat energy transferred to the air normally causes the air to increase in temperature. Hence, air exiting the [0093] heat exchanger 910 will have a temperature T_AirS1(Out), which is greater than T_AirS1(In). In instances where the air entering the heat exchanger is taken from an ambient source, such as, air near the structural passageway 912, heat energy transferred to the air at the heat exchanger 910 causes the air to expand and become less dense than the ambient air. In turn, the heated air rises in the structural passageway 912 according to principles of natural convection. Thus, natural convection currents enhance heat transfer at the heat exchanger 910.
A second stage of the two stage [0094] heat exchange system 900, Stage 2, includes a heat exchanger 920 to exchange heat between the working fluid and air. In Stage 2, an air duct 922 provides a flow path for air past or through the heat exchanger 920. In this example, a fan 924 consumes power 926 to propel air or cause air convection in the duct 922. Air convection caused by power 926 provided to the fan 924 aims to increase heat transfer between the working fluid and air at the heat exchanger 920.
According to [0095] Stage 2, the working fluid enters the heat exchanger 920 at a temperature T_WFS2(In) and air enters the heat exchanger 920 at a temperature T_AirS2(In) wherein T_WFS2(In) is greater than T_AirS2(In). At the heat exchanger 920, heat energy flows from the working fluid to the air. Depending on the state of the working fluid, the loss of heat energy normally causes the working fluid to experience at least a partial change of state (e.g., a phase change) and/or to decrease in temperature. For example, working fluid exiting the heat exchanger 920 may have a temperature T_WFS2(Out), which is less than T_WFS2(In).
Heat energy transferred to the air normally causes the air to increase in temperature. Hence, air exiting the [0096] heat exchanger 920 will have a temperature T_AirS2(Out), which is greater than T_AirS2(In). In instances where the air entering the heat exchanger is taken from an ambient source, heat energy transferred to the air at the heat exchanger 920 causes the air to expand and become less dense than the ambient air. In turn, the heated air may rise according to principles of natural convection. Thus, natural convection currents may enhance heat transfer at the heat exchanger 920. However, in this example, Stage 2 consumes power from the power source 926 to enhance heat exchange at the heat exchanger 920. In general, such power may be referred to as parasitic power. Parasitic power demands may depend on properties of the working fluid and/or working fluid flow requirements. For example, a requirement may be placed on T_WFS2(Out), which corresponds to a certain vapor/liquid phase composition of the working fluid at a given working fluid pressure.
While the exemplary two stage [0097] heat exchange system 900 shows a single heat exchanger 910 and a single structural passageway 912 in Stage 1, one or more heat exchangers and/or one or more structural passageways may be used in alternative examples. Further, while the exemplary two stage heat exchange system 900 shows a single heat exchanger 920, a single duct 922, a single fan 924 and a single power source 926 in Stage 2, one or more heat exchanges, one or more ducts, one or more fans, one or more power sources, and/or one or more structural passageways may be used in alternative examples. In general, Stage 1 relies predominantly on natural convection and Stage 2 relies, at least in part, on a power source.
Various exemplary heat exchange systems presented herein use a hot air tower or a structural passageway, optionally in conjunction with forced draft fans or the like, to reduce use of parasitic power. An exemplary system that combines use of a powerable convection unit (e.g., fan, etc.) and a tower or structural passageway, can reduce the required height of the tower or a structural passageway to a level that might be suitable for any application. [0098]
Various aforementioned examples consider an evaluation plant to demonstrate usefulness of an exemplary two stage heat exchange system. As described, a part-load tower may provide benefits at a reasonable cost that is considerably less than a conventional full-load tower. In one particular example, a part-load tower handles only a partial load and aims at achieving a maximum temperature for cooling air (e.g., considering constraints, etc.). Such a part-load tower optionally uses 22% of the airflow of a full-load tower. [0099]
In addition, an exemplary two stage heat exchange system approach is optionally suitable for implementation in conjunction with a structure, such as a building, wherein one or more structural passageways serve as natural convection routes that aid cooling. [0100]

Claims

1. A heat exchange system comprising:

a first stage including a first stage heat exchanger and a tower wherein

the first stage heat exchanger exchanges heat between a fluid and air to thereby heat the air and generate air convection in the tower; and

a second stage including a second stage heat exchanger and a powerable convection unit wherein the second stage heat exchanger receives the fluid from the first stage and exchanges heat between the fluid and air.

2. The heat exchange system of claim 1, wherein the fluid comprises a zeotrope.

3. The heat exchange system of claim 1, wherein at least some of the fluid enters the first stage as a vapor.

4. The heat exchange system of claim 1, wherein the fluid has a temperature glide for a vapor to liquid phase transition.

5. The heat exchange system of claim 1, wherein the air enters the first stage at ambient conditions.

6. The heat exchange system of claim 1, wherein the air exits the tower at a temperature greater than the ambient air temperature.

7. The heat exchange system of claim 1, wherein the fluid comprises ammonia and water.

8. The heat exchange system of claim 1, wherein the fluid exits a turbine prior to entering the first stage.

9. The heat exchange system of claim 1, wherein the first stage and the second stage condense the fluid as part of a Kalina cycle.

10. The heat exchange system of claim 1, wherein the first stage and the second stage condense the fluid as part of a cycle selected from the group consisting of Lorenz cycles, Uehara cycle, Rankine cycles, Carnot cycles and combinations thereof.

11. The heat exchange system of claim 1, wherein the tower has an air inlet located at least at the periphery of a cylindrical section.

12. The heat exchange system of claim 1, wherein the first stage heat exchanger includes a plurality of heat exchangers.

13. The heat exchange system of claim 1, wherein the first stage heat exchanger is located near an air inlet.

14. The heat exchange system of claim 1, wherein the first stage heat exchanger is located near a periphery of a cylindrical section of the tower.

15. The heat exchange system of claim 1, wherein the first stage includes a powerable convection unit.

16. The heat exchange system of claim 1, wherein the second stage includes a tower.

17. The heat exchange system of claim 1, wherein the height of the tower depends on one or more parameters associated with the second stage.

18. The heat exchange system of claim 1, wherein the height of the tower depends on a pinch point.

19. The heat exchange system of claim 1, wherein air enters the tower substantially horizontally and exits the tower substantially vertically.

20. The heat exchange system of claim 1, wherein the convection in the tower comprises natural convection.

21. A method comprising:

transferring heat energy from a fluid to air using a first heat exchanger to cause air convection in a tower; and

transferring heat energy from the fluid to air using a second heat exchanger and a fan.

22. The method of claim 21, wherein the fluid comprises a zeotrope.

23. The method of claim 21, wherein the fluid comprises ammonia and water.

24. The method of claim 21, wherein the first heat exchanger relies on the air convection in the tower.

25. The method of claim 21, wherein the fluid condenses from a vapor to a liquid.

26. The method of claim 21, wherein the fluid has a temperature glide for a vapor to liquid phase transition.

27. The method of claim 21, wherein a controller controls the fan.

28. The method of claim 21, wherein the air enters the first heat exchanger at ambient conditions.

29. The method of claim 21, wherein the air exits the tower at a temperature greater than the ambient air temperature.

30. The method of claim 21, wherein the height of the tower depends on a pinch point.

31. The method of 21, wherein the first heat exchanger and the second heat exchanger condense the fluid as part of a Kalina cycle.

32. The method of claim 21, wherein the first heat exchanger and the second heat exchanger condense the fluid as part of a cycle selected from the group consisting of Lorenz cycles, Uehara cycles, Rankine cycles, Carnot cycles and combinations thereof.

33. The method of claim 21, wherein the air convection in the tower comprises natural convection.

34. A heat exchange system comprising:

means for transferring heat energy from a fluid to air using only natural convection of the air; and

means for further transferring heat energy from the fluid to air using a fan.

35. The heat exchange system of claim 34, wherein the means for transferring heat energy comprises a heat exchanger.

36. The heat exchange system of claim 34, wherein the means for further transferring heat energy comprises a heat exchanger.

37. A heat exchange system comprising:

a first stage including a first stage heat exchanger and a structural passageway wherein the first stage heat exchanger exchanges heat between a fluid and air to thereby heat the air and generate air convection in the structural passageway; and

38. The heat exchange system of claim 37, wherein the structural passageway comprises a stairway.

39. The heat exchange system of claim 37, wherein the structural passageway is located at least partially within a building.

40. The heat exchange system of claim 37, wherein the air convection in the structural passageway comprises natural convection.