FIELD OF THE INVENTION

This invention relates to a thermal power plant. In particular, the invention relates to cooling of the working fluid of a binary geothermal power plant.

BACKGROUND OF THE INVENTION

Binary geothermal power plants are increasingly seen as a renewable energy source alternative to fossil fuel power plants and nuclear energy power plants. Binary geothermal power plants are safe to operate and are environmentally friendly.

Binary geothermal plants operate by pumping hot brine from a geothermal well to evaporate a working fluid via a heater heat exchanger. The evaporated working fluid is then passed through a turbine to generate electricity. Subsequently, the working fluid is cooled by a cooling heat exchanger to condense the working fluid. The working fluid is then pumped back to the heating heat exchanger to be evaporated again to complete a cycle. This cycle is closed-loop and the working fluid flows constantly to produce electricity.

In regions where water is available in sufficiently large quantities, water can be used as a heat sink to cool the working fluid via the cooling heat exchanger. However, many binary geothermal power plants are located in arid regions where ambient air is used as the heat sink to cool the working fluid via the cooling heat exchanger. Generally, as the ambient air temperature increases, the performance of air cooled binary type geothermal power plants decreases. The ambient air temperature may be such that the working fluid can not be cooled to within optimal design parameters for the power plant by air cooling.

OBJECT OF THE INVENTION

It is an object of the invention to overcome and/or alleviate one or more of the above disadvantages or to provide the consumer with a useful or commercial choice. SUMMARY OF THE INVENTION

In one form, although not necessarily the only or the broadest form, the invention resides in a cooling assembly for a thermal power plant, the cooling assembly including:

an ambient air cooling unit operable to decrease the temperature of a working fluid to a first temperature by heat exchange with ambient air; and a refrigeration unit operable to further decrease the temperature of the working fluid to a second temperature.

The refrigeration unit preferably includes a cooling heat exchanger in the form of an evaporator in which the working fluid is cooled by heat exchange with a coolant fluid of the refrigeration unit.

Preferably, the refrigeration unit operates in an absorption refrigeration cycle driven by a heat source.

Preferably, the heat source is solar radiation and the refrigeration unit includes a solar collector to extract heat from the solar radiation.

Preferably, the solar collector comprises an evacuated tube solar collector array.

Alternatively, the heat source is geothermal heat, combusting natural gas or an industrial waste heat stream.

The cooling assembly may further include a radiative cooling unit operable to decrease the temperature of the working fluid.

Preferably, the radiative cooling unit includes a thermal storage medium which is in heat transfer communication with the working fluid at a point after the temperature of the working fluid has decreased by air cooling, thereby to further decrease the temperature of the working fluid.

In one embodiment, the radiative cooling unit includes:

a radiative fluid;

a valve assembly operable to selectively connect the thermal storage medium to the solar collector to allow the radiative fluid to circulate between the solar collector and the thermal storage medium.

In one embodiment, the ambient air cooling unit is in the form of an air cooling tower. In accordance with another aspect of the invention there is provided a thermal power plant including the cooling assembly as defined and described hereinabove.

Preferably, the thermal power plant is connected to a geothermal reservoir.

Preferably, the thermal power plant operates on either a transcritical or supercritical thermodynamic cycle.

Preferably, the working fluid is carbon dioxide (CO ₂ ) based.

In accordance with yet another embodiment of the invention, there is provided a method of cooling a working fluid of a thermal power plant, the method including the steps of:

cooling the working fluid by ambient air cooling to a first temperature; and

thereafter further cooling the working fluid by refrigeration to a second temperature, which is lower than the first temperature.

Preferably, refrigeration includes operating a refrigeration unit in an absorption refrigeration cycle.

Preferably, the absorption refrigeration cycle is driven by heat from solar radiation.

Alternatively, the absorption refrigeration cycle is driven by geothermal heat, combusting natural gas or an industrial waste heat stream.

Preferably, the method includes the step of cooling the working fluid by heat exchange with a thermal storage medium which is cooled by radiating heat during the night.

Preferably, the working fluid is cooled by heat exchange with the thermal storage medium after the working fluid is cooled by ambient air cooling to the first temperature.

The step of refrigeration may include use of a solar collector to drive the solar refrigeration during the day, and the thermal storage medium is cooled during the night by circulating a radiative fluid between the solar collector and the thermal storage medium to cool the thermal storage medium by radiative cooling of the radiative fluid as it passes through the solar collector.

BRIEF DESCRIPTION OF THE DRAWINGS

An embodiment of the invention, by way of example only, will be described with reference to the accompanying drawings in which:

Figure 1 is a schematic layout of one embodiment of a binary geothermal power plant in accordance with the present invention;

Figure 2 is a temperature-entropy diagram for a transcritical thermodynamic cycle of a working fluid of the binary geothermal power plant of Figure 1 ;

Figure 3 is a temperature-entropy diagram for a supercritical thermodynamic cycle of a working fluid of the binary geothermal power plant of Figure 1 ;

Figure 4 is a schematic layout of the binary geothermal power plant of Figure 1 showing the components of the solar refrigeration unit of the power plant;

Figure 5 is a schematic layout of the binary geothermal power plant of Figure 1 further including a radiative cooling unit;

Figure 6 shows a schematic layout of another embodiment of a binary geothermal power plant in accordance with the invention, excluding a recuperator;

Figure 7 is a schematic flow diagram of the method of cooling the working fluid of a thermal power plant in accordance with the present invention;

Figure 8 is a schematic flow diagram of the method of Figure 7 including a further intermediate stage of cooling by thermal heat exchange with a thermal storage medium; and

Figure 9 is a graph showing the Monthly Electricity Generation by transcritical CO ₂ cycle, transcritical CO ₂ cycle including solar refrigeration cooling in accordance with one aspect of the invention, and steam Rankine cycle, simulated for the conditions of a typical Australian outback town. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

Figure 1 shows a schematic lay-out of one embodiment of a binary geothermal power plant 10. The power plant 10 includes a pump 12, a recuperator 14, a heater heat exchanger 16, a turbine 18, and a cooling assembly 20. A working fluid passes through the pump 12, recuperator 14, heater heat exchanger 16, turbine 18, and the cooling assembly 20. The path of the working fluid as it flows between the components is indicated by arrows.

The pump 12 is used to pressurize the working fluid from a low pressure to a high pressure. The pump 12 is connected to the recuperator 14 to pump working fluid through the recuperator 14.

The recuperator 14 is used to pre-heat the working fluid as it flows to the heater heat exchanger 16. The working fluid is pre-heated in the recuperator 14 by heat exchange with the working fluid exhausting from the turbine 18. The recuperator 14 is a heat exchanger that transfers some of the heat from the turbine 18 exhaust to preheat the working fluid. The recuperator 14 is connected to the heater heat exchanger 16 to flow preheated working fluid to the heater heat exchanger 16.

The heater heat exchanger 16 is used to heat the working fluid before it enters the turbine 18. The heater heat exchanger 16 is connected between the recuperator 14 and the turbine 18. A geothermal well 23 is also connected to the heater heat exchanger 16 to allow hot brine from the well 23 to flow through the heater heat exchanger 16. The working fluid is heated by heat exchange in the heater heat exchanger 16 with the hot brine from the geothermal well 23.

The turbine 18 is used to drive an electricity generator (not shown). The working fluid expands from high pressure to low pressure as it drives the turbine 18. The turbine 18 is connected to the cooling assembly 20 via the recuperator 14.

The cooling assembly 20 is used to cool the working fluid. The cooling assembly 20 includes an ambient air cooling unit in the form of an air cooling tower 22 and a solar refrigeration unit 24. The air cooling tower 22 is formed from a tower component and heat exchanger tubes. The heat exchanger tubes are arranged about the base of the tower component. In use, the working fluid circulates through the heat exchanger tubes, heating the air around the tubes and being cooled as it does so by heat exchange. The heated air rises through the tower component and cooler ambient air is sucked past the heat exchanger tubes to replace the air as it rises.

The solar refrigeration unit 24 includes a solar collector to drive the solar refrigeration in an absorption refrigeration cycle during the day. The solar refrigeration unit 24 is described in more detail with reference to Figure 4.

The working fluid is CO ₂ . The abrupt property changes for CO ₂ near its critical temperature, which is at about 30 ⁰ C, is suitable for its use with air cooled geothermal plants.

Figure 2 shows a temperature-entropy diagram for a thermodynamic cycle (known as a transcritical cycle) for the working fluid as it passes through the components of the power plant 10. The different states of the working fluid along the transcritical cycle are shown as state points in Figure 2, which cross references to points in the power plant 10 of Figure 1. A curve 11 in Figure 2 represents the phase diagram for the working fluid.

Further referring to Figures 1 and 2, the transcritical thermodynamic cycle of the working fluid starts with the working fluid at the saturated liquid phase at point 1 on exit from the cooling assembly 20. The working fluid is pumped to point 2 by the pump 12 and is heated in the recuperator 14 to point 3 using the residual heat in the working fluid exhausting from the turbine 18. Heating of the working fluid from 3 to 4 is by heat exchange with hot brine from the well 23 flowing through the heater heat exchanger 16. At point 4, the working fluid is expanded through the turbine 18 to generate power which is used to drive an electricity generator. The turbine 18 exit state of the working fluid is denoted by 5. From 5 to 6, the working fluid exiting from the turbine 18 exchanges some of its residual heat to the working fluid flowing through the recuperator 14 to the heater heat exchanger 16. The recuperator 14 exit point for the working fluid is 6. The working fluid is then cooled by ambient temperature air in the air cooling tower 22 to a first temperature at 7.

The extent of air cooling in the air cooling tower 22 depends on the ambient air temperature at that time. The cycle is designed for a particular condensation temperature, called the design-point condensation temperature. Choice of the design-point condensation temperature is dependant on the local weather. If the ambient air temperature is low enough to enable cooling of the working fluid in the air cooling tower 22 to the design-point condensation temperature, then no further cooling of the working fluid by the solar refrigeration unit 24 is required. If the ambient air temperature is higher than what is required to cool the working fluid to the design-point condensation temperature, then after exit from the air cooling tower 22 the working fluid is further cooled to a second temperature (the design-point condensation temperature) at 1 through solar refrigeration by the solar refrigeration unit 24. The second temperature at 1 is lower than the first temperature at 7.

Figure 3 shows a temperature-entropy diagram for a thermodynamic cycle (known as a supercritical cycle) for the working fluid as it passes through the components of the power plant 10. The lay-out for a geothermal power plant 10 based on a supercritical cycle and using both air cooling and solar refrigeration is the same as shown in Figure 1. The supercritical thermodynamic cycle is used when the cooling assembly 20 is unable to condense the working fluid. A curve 11 in Figure 3 represents the phase diagram for the working fluid. A peak "a" of the curve 11 corresponds to a critical condenser temperature. Condensation is possible only below the critical condenser temperature. If the prevailing ambient air temperature is so high that the cooling unit 20 can not cool the working fluid below the critical condenser temperature through most of the year, it may then be appropriate to use a supercritical cycle where no condensation of the working fluid takes place.

Further referring to Figures 1 and 3, the supercritical cycle starts with the working fluid at a supercritical state at point 1 on exit from the cooling assembly 20. The working fluid is compressed to point 2 by the pump 12 working as a compressor and is heated in the recuperator 14 to point 3 using the residual heat in the working fluid exiting from the turbine 18. Heating of the working fluid from 3 to 4 is done by heat exchange with the hot brine flowing through the heater heat exchanger 16. At point 4, the working fluid is expanded through the turbine 18 and generates power. The turbine 18 exit state of the working fluid is denoted by 5. From 5 to 6, the working fluid exiting from the turbine 18 exchanges some of its residual heat with the working fluid flowing through the recuperator 14 on the way to the heater heat exchanger 16. The recuperator 14 exit point for the hot working fluid is 6. The working fluid is then cooled by air to 7 in the air cooling tower 22.

For the same geothermal resource, a transcritical cycle delivers a higher efficiency than a supercritical cycle provided the working fluid can be cooled enough to be condensed. Under some conditions, it may not be economically feasible to provide additional cooling to temperatures to condense the working fluid. In such instances a supercritical cycle is used. The extent of air cooling in the air cooling tower depends on the ambient air temperature at that time. Similar to a transcritical cycle, a supercritical cycle is also designed for a certain minimum temperature value. This corresponds to the minimum temperature in the cycle and called the design-point minimum temperature. If the ambient air temperature is low enough to enable cooling of the working fluid in the air cooling tower 22 to a temperature below its design-point minimum temperature for the supercritical cycle, then no further cooling of the working fluid by the solar refrigeration unit 24 is required. If the ambient air temperature is higher than what is required to cool the working fluid to the design-point minimum temperature, then after exit from the air cooling tower 22 the working fluid is further cooled to a second temperature (the design-point minimum temperature) at 1 through solar refrigeration by the solar refrigeration unit 24. The second temperature at 1 is lower than the first temperature at 7.

Figure 4 shows the power plant 10 and specifically the components of the solar refrigeration unit 24. The solar refrigeration unit 24 includes a solar collector 26 to drive the solar refrigeration in an absorption refrigeration cycle during the day. The solar collector 26 extracts heat from solar radiation as a heat source to drive the absorption refrigeration cycle of the solar refrigeration unit 24. The solar collector 26 is in the form of an evacuated tube solar collector array. The solar refrigeration unit 24 further includes a condenser 28, a coolant restrictor valve 30, an evaporator 32, an absorber 34, a solution pump 36, a solution throttle valve 38 and a generator 40. A coolant fluid flows in the solar refrigeration unit 24. The solar collector 26 is connected to the generator 40 as a heat source for the absorption refrigeration cycle.

Working fluid exiting the cooling tower 22 is cooled in the evaporator 32 by heat exchange with the coolant fluid. The coolant fluid is evaporated in the evaporator 32 by the heat exchange with the working fluid. The evaporator 32 is a cooling heat exchanger where the working fluid is cooled to the second temperature. The vaporized coolant fluid flows from the evaporator 32 into the absorber 34. The coolant fluid is absorbed in the absorber 34 by a solvent, for example NH3 or LiBr, to form a coolant solution.

The coolant solution is then pumped to the generator 40 using the solution pump 36. At the generator 40, solar heat is applied to the coolant solution by heat exchange with a collector fluid which circulates between the generator 40 and the solar collector 26. Heating the coolant solution separates the coolant fluid from the solvent. The separated solvent is depressurised in the solution throttle valve 38 to the pressure of the absorber 34. The coolant fluid flows from the generator 40 to the condenser 28, where the coolant fluid is condensed. The condensed coolant fluid then flows back to the evaporator 32 via the coolant restrictor valve 38 to repeat the cycle to continuously cool the working fluid in the power plant 10.

Although absorption refrigeration has been described with reference to solar radiation being the heat source via the solar collector 26, the Applicant envisages that absorption refrigeration for the power plant may alternatively be driven by heat sources such as geothermal heat from the well 23, combusting natural gas or an industrial waste heat stream. These alternative heat sources will thus heat the coolant solution in the generator 40.

Figure 5 shows the power plant 10 with an optional intermediate stage of cooling of the working fluid between air cooling at the air cooling tower 22 and solar refrigeration cooling at the solar refrigeration unit 24. The intermediate stage of cooling is by a radiative cooling unit 42 having a thermal storage medium 44. The thermal storage medium 44 is cooled during the night by heat exchange with a radiative fluid circulated through the solar collector 26 of the solar refrigeration unit 24. The radiative cooling unit 42 cools the working fluid during day by heat exchange with the thermal storage medium 44. The radiative cooling unit 42 comprises the thermal storage medium 44, a working fluid heat exchanger 46, a pump 48, a valve assembly 50, a radiative fluid heat exchanger 52 and the radiative fluid flowing between the radiative fluid heat exchanger 52 and the solar collector 26.

The pump 48 is used to circulate the thermal storage medium through the working fluid heat exchanger 46 during the day.

The valve assembly 50 is used to selectively fluidly connect the radiative fluid heat exchanger 52 to the solar collector 26 during the night. The radiative fluid thus circulates between the solar collector 26 and the radiative fluid heat exchanger 52 during night when the solar collector 26 is not in use as part of the solar refrigeration unit 24.

The solar collector 26 is used to pass the radiative fluid therethrough to cool the radiative fluid by thermal radiation during the night. The same solar collector 26 may be used during the day to drive the solar refrigeration unit 24. The valve assembly 50 is operable between a condition wherein it opens a flow path for radiative fluid to circulate between the solar collector 26 and the radiative fluid heat exchanger 52, and a condition wherein it opens a flow path for a collector fluid to circulate between the generator 40 and the solar collector 26.

The radiative fluid cools the thermal storage medium 44 via the radiative fluid heat exchanger 52 as the radiative fluid is cooled by radiative cooling in the solar collector 26.

Radiative cooling at night of a thermal storage medium is practicable for most arid regions as these regions enjoy nights with very clear skies and low sky temperatures, which makes radiative cooling effective. The thermal storage medium 44 is used to cool the working fluid during the day by heat exchange to an intermediate temperature between the first temperature at 7 and the second temperature at 1. Cooling by the radiative cooling unit 42 reduces reliance on solar refrigeration.

Figure 6 shows a schematic layout of another embodiment of a binary geothermal plant 100. The recuperator 14 is optional in the power plant 10 for both transcritical and supercritical cycles. The only difference between the power plant 10 and the power plant 100 is the exclusion of a recuperator from the power plant 100. The pump 12 of the power plant 100 is directly connected to the heater heat exchanger 16. In Figure 6, components of the power plant 100 which are the same as the components of the power plant

10 are indicated by the same reference numerals.

Figure 7 shows a diagram of the method 200 of cooling the working fluid described with reference to the thermal power plants 10, 100 above.

The working fluid of the thermal power plants 10, 100 is first cooled by ambient air cooling at step 202 to a first temperature. Ambient air cooling of the working fluid is by operation of the cooling tower 22 as described with reference to Figures 1-6.

The working fluid is further cooled by refrigeration at step 204 to a second temperature, lower then the first temperature. Cooling of the working fluid by refrigeration includes absorption refrigeration 206. Absorption refrigeration is driven by heat from a heat source 208. The heat source 208 is either solar radiation 210, geothermal heat 212, combusting natural gas

214 or an industrial waste heat stream 216. A coolant fluid 218 cooled during absorption refrigeration 206 cools the working fluid 220 to the second temperature by heat exchange 222 with the working fluid 220.

Figure 8 shows a diagram of the method 200 further including an intermediate stage 230 of cooling the working fluid 232 by thermal heat exchange 234 with a thermal storage medium 236. The working fluid 232 is the same working fluid 220 referred to in step 204. The thermal storage medium 236 is cooled at night by heat radiation 238 of a radiative fluid. The thermal storage medium 236 is cooled during the night by heat exchange with the radiative fluid as described with reference to Figure 5.

Referring to Figure 9, the Monthly Electricity Generation of:

a) a transcritical CO ₂ binary geothermal power plant without a solar refrigeration unit;

b) a transcritical CO ₂ binary geothermal power plant with a solar refrigeration unit ; and

c) a steam Rankine cycle geothermal power plant;

is simulated over a year. Design parameters were for a hot brine temperature of 250 ⁰ C and a flow rate of 500 kg/s. The hot brine temperature and the flow rate are representative of hot fractured rock geothermal prospects currently being explored in Central Australia. The brine temperature of 25O ⁰ C is too hot for organic fluids such as isopentane and a steam cycle is the only conventional option. The solar refrigeration unit design parameters included:

Solar collector fluid temperature = 160 ⁰ C

Coefficient of performance for absorption refrigeration = 1.0

Solar collection efficiency = 0.80

Since the ambient temperature varies daily and seasonally, the performance of an air-cooled plant geothermal power plant varies through the year. Air-cooled geothermal plants are designed to perform optimally (the design point performance) at a temperature of the working fluid after cooling (the design point temperature). The performance drop at higher ambient temperatures is severe for the transcritical CO ₂ cycles shown in Figure 9 because the CO ₂ pump power consumption increases rapidly as the working fluid temperature after cooling moves away from the design point temperature. From the graph it is apparent that while the design-point performance for the transcritical CO ₂ cycles is relatively better than for the steam Rankine cycle, their off-design point performance is very poor. The steam Rankine plant power output on the other hand is shown to be reasonably steady through the year and while its performance would diminish on hot days, the reduction is relatively small. The Applicant envisages that the addition of solar refrigeration unit may make the transcritical CO ₂ cycle outperform the steam Rankine cycle for the given conditions.

While the off-design performance of an air-cooled binary geothermal power plant is poor on hot days, the nature of the transcritical and supercritical cycles make it possible to provide cooling in stages. The present invention is particularly suited for hot arid regions as a first stage of cooling can be provided by ambient air cooling and since performance drop of the power plant occurs at the same time when the solar radiation is at its maximum, the second stage of cooling is provided by using a solar refrigeration unit to maintain the design heat sink conditions for the power plant. The Applicant envisages that the extra power generated by the power plants 10, 100 because of the inclusion of the solar refrigeration unit 24 may be substantially higher than the electricity that could be generated if the solar collector of the solar refrigeration unit 22 was driving a solar thermal power plant separate from the power plants 10, 100.

It will also be appreciated that various other changes and modifications may be made to the embodiment described without departing from the spirit and scope of the invention. For example, solar refrigeration may be achieved by photovoltaic operated refrigeration cycle or solar mechanical refrigeration.