Patent Grant
11466907
The subject matter herein generally relates to turbo-compression cooling. More specifically, the subject matter herein relates to a system implementing a turbine coupled with a compressor to utilize low-grade waste heat to power a cooling cycle in various applications.
Industries of all types produce waste heat because of inefficient system performance. Conventionally, the waste heat is exhausted into the environment in the form of steam, heated water or air, or hot exhaust gases. As an example, an industrial plant operating a steam generator produces waste heat that is exhausted into the environment through its exhaust stack. As another example, marine diesel engines include various thermal energy streams (e.g., engine jacket water, lubrication oil, and aftercooler air) that reject heat to seawater. These systems, among others, can increase overall efficiency by recovering waste heat. Current systems for waste heat recovery, however, are large, cumbersome, and expensive. Additionally, each industry includes its own unique challenges to utilizing waste heat to improve system performance. It is with these thoughts in mind, among others, that aspects of the ultra-efficient turbo-compression cooling systems were developed.
A turbo-compression cooling system includes a power cycle and a cooling cycle coupled one to the other. The power cycle implementing a waste heat boiler configured to evaporate a first working fluid and a turbine configured to receive the evaporated working fluid. The turbine having a plurality of vanes disposed around a central shaft and configured to rotate as the first working fluid expands to a lower pressure within the turbine. A condenser then condenses the first working fluid to a saturated or subcooled liquid and a mechanical pump pumps the liquid to reenter the waste heat boiler. The cooling cycle implements a compressor configured to increase the pressure of a second working fluid, a condenser configured to condense the second working fluid to a saturated or subcooled liquid upon exiting the compressor, an expansion valve wherein the second working fluid expands to a lower pressure, and an evaporator rejecting heat from a circulating fluid to the second working fluid, thereby cooling the circulating fluid. The turbine and compressor can be coupled one to the other, thereby coupling the power cycle and the cooling cycle.
Aspects of the present disclosure may include a system for turbo-compression cooling. In certain instances, the system is a shipboard system on a marine vessel. The shipboard system may include a power cycle and a cooling cycle. The power cycle may include a first working fluid, a waste heat boiler configured to evaporate the working fluid, a turbine, and a condenser. The turbine receiving the evaporated working fluid, and having a plurality of vanes disposed around a central shaft and configured to rotate about the central shaft. The plurality of vanes are configured to rotate as the working fluid expands to a lower pressure. The condenser condenses the working fluid to a saturated or subcooled liquid. The cooling cycle may include a second working fluid, a first compressor configured to increase the pressure of the second working fluid, a condenser configured to condense the second working fluid to a saturated or subcooled liquid after exiting the first compressor, an expansion valve, and an evaporator. The expansion valve configured to expand the second working fluid to a lower pressure. The evaporator rejecting heat from a circulating fluid to the second working fluid, thereby cooling the circulating fluid. The turbine and first compressor are coupled one to the other, thereby coupling the power cycle and the cooling cycle. The waste heat boiler receives waste heat from engine jacket water and lubricating oil from a ship service diesel generator. In addition, the evaporator cools water in a shipboard cooling loop.
In certain instances, the water in the shipboard cooling loop may be seawater that cools shipboard chillers in the marine vessel. In certain instances, the shipboard cooling loop provides cooling within the marine vessel.
In certain instances, the cooling cycle further may include a second compressor configured to discharge the second working fluid to the first compressor. In certain instances, the second compressor may be electrically powered.
In certain instances, the first compressor may be powered via the waste heat from the waste heat boiler.
In certain instances, the first working fluid and the second working fluid are the same fluid.
In certain instances, the power cycle and first working fluid are hermetically sealed from the cooling cycle and the second working fluid.
In certain instances, the turbine and the first compressor are magnetically coupled one to the other, and the turbine has a first shaft and the first compressor has a second shaft, one of the first shaft and the second shaft disposed around at least a portion of the other of the first shaft and the second shaft, the first shaft having one or more first polarity magnetic elements and the second shaft having one or more second polarity magnetic elements, the first polarity and the second polarity being opposite and magnetically engaged with one another.
In certain instances, the system further may include a recuperator configured to receive heat rejected by the first working fluid, and the recuperator transfers the rejected heat to the saturated or subcooled liquid as the working fluid re-enters the waste heat boiler.
Aspects of the present disclosure may include a method of turbo-compression cooling. The method may include the following steps. The method may include receiving, from a waste heat source, heat waste in a waste heat boiler. The method may include evaporating a first working fluid using the heat waste in the waste heat boiler. The method may include generating mechanical power through expansion of the first working fluid to a lower pressure in a turbine, the expansion of the first working fluid rotating one or more turbine vanes. The method may include condensing the first working fluid to a saturated or subcooled liquid in a condenser. The method may include pressurizing the saturated or subcooled liquid through a mechanical pump to re-enter the waste heat boiler. The method may include transferring the generated mechanical power to a first compressor, the compressor configured to receive a second working fluid. The method may include compressing the second working fluid via the first compressor thereby increasing the pressure of the vapor. The method may include condensing the second working fluid in a condenser to a saturated or subcooled liquid. The method may include expanding the second working fluid to a lower pressure in an expansion valve. In addition, the method may include rejecting heat through an evaporator from circulating cooling fluid to the second working fluid.
In certain instances, the method further may include compressing the second working fluid via a second compressor powered separately from the first compressor; and discharging the second working fluid from the second compressor to the first compressor. In certain instances, the first working fluid and the second working fluid are the same fluid.
In certain instances, the method further may include rejecting heat from the first working fluid exiting the turbine in a recuperator, and absorbing heat in the first working fluid exiting the mechanical pump.
In certain instances, the method further may include rejecting heat from the second working fluid in the recuperator.
In certain instances, the method further may include recirculating a portion of the second working fluid exiting the first compressor to an inflow of the second compressor to bypass the recuperator.
In certain instances, the method further may include preheating the second working fluid in a suction-line heat exchanger prior to compressing the second working fluid via the second compressor.
In certain instances, the second working fluid may be preheated via the second working fluid after discharge from the condenser.
In certain instances, the method further may include recirculating a portion of the second working fluid exiting the first compressor to an inflow of the first compressor to bypass the recuperator.
In certain instances, the power cycle and first working fluid are hermetically sealed from the cooling cycle and the second working fluid.
In certain instances, the second compressor may be electrically powered.
In certain instances, the method further may include rejecting heat from the second working fluid exiting the second compressor in an economizer, and absorbing heat in the first working fluid exiting the mechanical pump. In certain instances, the method may further include rejecting heat from the first working fluid in an economizer.
In certain instances, the method further may include rejecting heat from the first working fluid exiting the turbine in a recuperator, and absorbing heat in the first working fluid exiting the mechanical pump.
Aspects of the present disclosure may include a system for turbo-compression cooling in a facility with waste heat. The system may include a power cycle and a cooling cycle. The power cycle may include a first working fluid, a waste heat boiler configured to evaporate the working fluid, a turbine receiving the evaporated working fluid, the turbine having a plurality of vanes disposed around a central shaft and configured to rotate about the central shaft, the plurality of vanes configured to rotate as the working fluid expanding to a lower pressure, and a condenser condensing the working fluid to a saturated or subcooled liquid. The cooling cycle may include a second working fluid, a first compressor configured to increase the pressure of the second working fluid, a condenser configured to condense the second working fluid to a saturated or subcooled liquid after exiting the first compressor, an expansion valve configured to expand the second fluid to a lower pressure upon passing therethrough, and an evaporator rejecting heat from a circulating fluid to the second working fluid, thereby cooling the circulating fluid. The turbine and first compressor are coupled one to the other, thereby coupling the power cycle and the cooling cycle. The waste heat boiler may be configured to receive the waste heat from the facility, which has a centralized refrigeration system, which can have ammonia as the working fluid. The evaporator in the turbo-compression cooling system precools ammonia in the refrigeration system to lower the amount of compressor work required for the refrigeration system.
In certain instances, the ammonia may be cooled from about 27° C. to about 6.5° C.
In certain instances, the ammonia may be precooled between a receiver of the ammonia-cooling loop and an accumulator of the ammonia-cooling loop.
In certain instances, the ammonia may be precooled between the receiver of the ammonia-cooling loop and an expansion valve of the ammonia-cooling loop.
In certain instances, the ammonia may be precooled between a cold box of the ammonia-cooling loop and an accumulator of the ammonia-cooling loop.
Aspects of the present disclosure may include a system for turbo-compression cooling in a distributed refrigeration system in a facility producing waste heat. The distributed refrigeration system may include an accumulator receiving a two-phase mixture of a refrigerant from an expansion valve. The refrigerant in a liquid state from the accumulator may be repressurized and sent to a distributed set of evaporators. The system may include a power cycle and a cooling cycle. The power cycle may include a power cycle fluid, a waste heat boiler configured to evaporate the power cycle fluid, the waste heat boiler receiving waste heat from the facility, a turbine, and a condenser. The turbine receives the evaporated power cycle fluid. The turbine includes a plurality of vanes disposed around a central shaft and configured to rotate about the central shaft. The plurality of vanes configured to rotate as the power cycle fluid expands to a lower pressure. The condenser condenses the power cycle fluid to a saturated liquid. The cooling cycle may include a cooling cycle fluid, a first compressor configured to increase the pressure of the cooling cycle fluid. The first compressor may be coupled to the turbine, thereby coupling the power cycle and the cooling cycle. The cooling cycle further includes a condenser configured to condense the cooling cycle fluid to a saturated liquid after exiting the first compressor; an expansion valve configured to expand the cooling cycle fluid to a lower pressure, and an evaporator integrated with the refrigerant of the distributed refrigeration system so as to provide cooling to the refrigerant, the refrigerant rejecting heat to the second working fluid.
In certain instances, the refrigerant may be cooled prior to entering the expansion valve.
In certain instances, the refrigerant may be cooled after exiting a receiver of the distributed refrigeration system.
In certain instances, the refrigerant may be ammonia.
In certain instances, the refrigerant may be cooled from about 27 degrees Celsius to about 6.5 degrees Celsius.
In certain instances, the refrigerant may be cooled about 20.5 degrees Celsius.
In certain instances, cooling of the refrigerant via the evaporator of the cooling cycle decreases vapor percentage of the two-phase mixture of the refrigerant received in the accumulator.
Aspects of the present disclosure may include a method of turbo-compression cooling. The method may include the following steps. Receiving, from a waste heat source, heat waste in a waste heat boiler. Evaporating a first working fluid using the heat waste in the waste heat boiler. Generating mechanical power through expansion of the first working fluid to a lower pressure in a turbine, the expansion of the first working fluid rotating one or more turbine vanes. Condensing the first working fluid to a saturated or subcooled liquid in a condenser. Pressurizing the saturated or subcooled liquid through a mechanical pump to re-enter the waste heat boiler. Transferring the generated mechanical power to a compressor, the compressor configured to receive a second working fluid. Compressing the second working fluid via the compressor thereby increasing the pressure of the saturated vapor. Condensing the second working fluid in a condenser to a saturated or subcooled liquid. Expanding the second working fluid to a lower pressure via an expansion valve. Rejecting heat through an evaporator from circulating cooling fluid to the second working fluid. In addition, rejecting heat from the second working fluid exiting the compressor in an economizer, and absorbing heat in the first working fluid exiting the mechanical pump in the economizer.
In certain instances, the first working fluid and the second working fluid are the same fluid.
In certain instances the method may further rejecting heat from the first working fluid exiting the turbine in a recuperator, and absorbing heat in the first working fluid exiting the mechanical pump in the recuperator.
In certain instances the method may further rejecting heat from the second working fluid in the recuperator.
In certain instances the method may further rejecting heat from the second working fluid in a recuperator.
In certain instances the method may further recirculating a portion of the second working fluid exiting the compressor to bypass the recuperator.
In certain instances the method may further preheating the second working fluid in a suction-line heat exchanger prior to compressing the second working fluid via the compressor.
In certain instances, the power cycle and first working fluid are hermetically sealed from the cooling cycle and the second working fluid.
Aspects of the present disclosure may include a method of turbo-compression cooling. The method may include the following steps. Receiving, from a waste heat source, heat waste in a waste heat boiler. Evaporating a first working fluid using the heat waste in the waste heat boiler. Generating mechanical power through expansion of the first working fluid to a lower pressure in a turbine, the expansion of the first working fluid rotating one or more turbine vanes. Condensing the first working fluid to a saturated or subcooled liquid in a condenser. Pressurizing the saturated or subcooled liquid through a mechanical pump to re-enter the waste heat boiler. Transferring the generated mechanical power to a compressor, the compressor configured to receive a second working fluid. Compressing the second working fluid via the compressor thereby increasing the pressure of the saturated vapor. Condensing the second working fluid in a condenser to a saturated or subcooled liquid. Expanding the second working fluid to a lower pressure via an expansion valve. Rejecting heat through an evaporator from circulating cooling fluid to the second working fluid. In addition, preheating the second working fluid in a suction-line heat exchanger prior to compressing the second working fluid via the compressor.
In certain instances, the first working fluid and the second working fluid are the same fluid.
In certain instances the method may further rejecting heat from the first working fluid exiting the turbine in a recuperator, and absorbing heat in the first working fluid exiting the mechanical pump in the recuperator.
In certain instances the method may further rejecting heat from the second working fluid in the recuperator.
In certain instances the method may further rejecting heat from the second working fluid in a recuperator.
In certain instances, the method may further include recirculating a portion of the second working fluid exiting the compressor to bypass the recuperator.
In certain instances, the second working fluid is preheated in the suction-line heat exchanger via the second working fluid exiting the condenser.
In certain instances, the power cycle and first working fluid are hermetically sealed from the cooling cycle and the second working fluid.
Implementations of the present technology will now be described, by way of example only, with reference to the attached figures, wherein:
It will be appreciated that for simplicity and clarity of illustration, where appropriate, reference numerals have been repeated among the different figures to indicate corresponding or analogous elements. In addition, numerous specific details are set forth in order to provide a thorough understanding of the embodiments described herein. However, it will be understood by those of ordinary skill in the art that the embodiments described herein can be practiced without these specific details. In other instances, methods, procedures and components have not been described in detail so as not to obscure the related relevant feature being described. The drawings are not necessarily to scale and the proportions of certain parts may be exaggerated to better illustrate details and features. The description is not to be considered as limiting the scope of the embodiments described herein.
Several definitions that apply throughout this disclosure will now be presented.
The term “coupled” is defined as connected, whether directly or indirectly through intervening components, and is not necessarily limited to physical connections. The connection can be such that the objects are permanently connected or releasably connected. The term “substantially” is defined to be essentially conforming to the particular dimension, shape or other word that substantially modifies, such that the component need not be exact. For example, substantially cylindrical means that the object resembles a cylinder, but can have one or more deviations from a true cylinder. The term “comprising” means, “including, but not necessarily limited to”; it specifically indicates open-ended inclusion or membership in a so-described combination, group, series and the like.
A “thermal fluid” is defined as any working fluid optimized for use in a power/heating cycle. A “cooling fluid” is defined as any working fluid optimized for use in a cooling/refrigeration cycle. In some instances, a thermal fluid and cooling fluid can be the same, such as water, which can operate both a power cycle and cooling cycle.
The following portion of the application is broken into three sections describing different implementations of a turbo-compression cooling system. Section I describes using a turbo-compression cooling system in an ammonia-cooling loop. Section II describes using a turbo-compression cooling system in a marine diesel generation set. In addition, Section III describes a turbo-compression cooling system in a generalized industrial context. While Sections I and II describe specific implementations and context for the use of a turbo-compression cooling system, the systems outlined in those sections are applicable for use in other systems that generate waste heat. For example, Section II describes utilizing waste heat from engine jacket water and lubricating oil. Other generator systems may generate the same or similar low-grade waste heat that could benefit from the turbo-compression cooling system; thus, the description in the sections are not limiting. Similarly, the turbo-compression cooling system described in Section I is applicable to other systems with similar principles. All descriptions of the turbo-compression cooling system utilize ultra-low temperature waste heat in order to power the power cycle (organic or inorganic Rankine cycle) portion system, which in turn drives a compressor in the cooling cycle (refrigeration cycle). Therefore, the turbo-compression cooling systems as described herein may be applied to any system that generates low temperature waste heat and desires efficiency gains in the overall system. Additionally, the components of the individually described systems may be applied to the other systems without limitation.
I. Distributed Cooling Systems and Turbo-Compression Cooling System
The cooling system 100 includes a compressor 102 that compresses the fluid (e.g., ammonia) 104, which causes the fluid 104 to heat up as it is pressurized. The fluid 104 is received within a cooling tower 106 where heat from the fluid 104 dissipates. As the fluid 104 dissipates heat, the fluid 104 condenses and converts to a liquid, still at high pressure. A receiver 108 receives the fluid 104. In certain instances, the fluid 104 is received by the receiver 108 at a pressure of about 140 pounds per square inch gauge (PSIG). The receiver 108 is a storage vessel for the fluid 104 that is designed to hold excess fluid 104 not in circulation. The receiver 108 may contain a filter for trapping debris within the line. After the receiver 108, the fluid 104 travels through an expansion valve 110. The expansion valve 110 causes a decrease in pressure in the fluid 104 by, for example, passing it through a small diameter tube or orifice. In the context of an ammonia cooling system 100, the fluid 104 may enter the expansion valve 110 at about +27° C., and, upon exiting the expansion valve 110, the fluid 104 begins to boil at −9° C.
The fluid 104 is then received in an accumulator 112 at a pressure of about 32 PSIG. The accumulator 112 is a filter for the system 100 that includes a desiccant. The accumulator 112 prevents liquid fluid 104 from entering the compressor 102, which only pumps vapor, not liquid. The accumulator 112 has two outlets, a vapor outlet 114 and a liquid outlet 116. At the vapor outlet 114 is a metering ejector that vaporizes the fluid 104 and sends it to the compressor 102. In the context of an ammonia cooling system 100, about 13.6% of ammonia is routed from the accumulator 112, out the vapor outlet 114 (i.e., fluid 104 is 100% vapor), and directly to the compressor 102. Stated differently, 13.6% of the fluid 104 (i.e., the vapor) is not utilized for cooling, but is rerouted through the system 100.
The fluid 104 exiting the liquid outlet 116 is received by a pump 118 that pumps the fluid 104 to about 60 PSIG. The fluid 104 is then received by an expansion valve 120, which is located very near the evaporator within the cold box 122. After the expansion valve 120, the fluid 104 then enters a refrigerated space or cold box 122 where the pressure is about 32 PSIG. The refrigerated space 122 may be used within the industrial facility to cool a product (e.g., food, beverages). The cold box 122 may contain one or more evaporators therein. The fluid 104 is evaporated in the refrigerated space 122 as heat from the product heats the fluid 104. Upon exiting the refrigerated space 122, the fluid 104 mixes with the vapor fluid 104 exiting the vapor outlet 114 of the accumulator 112 at a valve 124. The combined fluid 104 then enters into the compressor 102 to complete the loop of the cooling system 100.
The efficiency of the cooling system 100 can be increased by decreasing the percentage of vapor fluid 104 that bypasses the refrigerated space 122 by being recirculated through the system 100. Stated differently, the efficiency of the cooling system 100 can be increased by increasing the percentage of liquid fluid 104 in the accumulator 112. This can be done by decreasing the temperature of the incoming fluid 104 into the expansion valve 110. Accordingly, a turbo compression cooling system 200, as shown in
The power cycle 202 operates with a first working fluid 208 receiving waste heat from a waste heat source, such as the heat dissipated in the cooling tower 106 of the cooling system 100 of
The turbine 206 can have a plurality of vanes coupled to a shaft 270, the plurality of vanes configured to impart rotation upon the shaft 270 as the first working fluid 208 expands within the turbine 206. Expansion of the first working fluid 208 within the turbine 206 generates mechanical power, thus rotating the shaft 270.
In some instances, the turbine 206 can be a multi-stage turbine having a plurality of vanes arranged to allow expansion of the first working fluid 208 and a second plurality of vanes arranged to allow further expansion of the first working fluid 208. The plurality of vanes and the plurality of second vanes are arranged for optimal performance based on the operating pressures, temperatures, and first working fluid 208 of the power cycle 202 of the ultra-efficient turbo-compressor cooling system 200. For a further description of the turbo compression cooling system, among other subject matter, reference is made to U.S. Pat. No. 10,294,826, which is hereby incorporated by reference in its entirety.
Upon exiting the turbine 206, the first working fluid 208 enters a condenser 216. In certain instances, the condenser 216 may be a dry air condenser, or a wet air condenser. The condenser 216 condenses the first working fluid 208 from a vapor to a saturated liquid. The condenser 216 can be an air-cooled heat exchanger allowing the first working fluid 208 to reject heat to the environment. The first working fluid 208 leaves the condenser 216 as a saturated liquid and enters a mechanical pump 218. The mechanical pump 218 re-pressurizes the first working fluid 208 and circulates the working fluid 208 back to the waste heat boiler 210.
While the ultra-efficient turbo-compressor cooling system 200 is shown and described with respect to the power cycle 202 as shown in
Still referring to
In some instances, the compressor 252 can be a multi-stage compressor having a plurality of impellers arranged to allow compression of the second working fluid 254 and a second plurality of impellers arranged to allow further expansion of the second working fluid 254. The plurality of impellers and the plurality of second impellers are arranged for optimal performance based on the operating pressures, temperatures, and second working fluid 208 of the cooling cycle 250 of the ultra-efficient turbo-compressor cooling system 200.
The condenser 256 is an air-cooled heat exchanger condensing the second working fluid 254 from a slightly superheated vapor to a saturated or subcooled liquid. The condenser 256 can have a forced airflow across the heat exchanger to increase efficiency and cooling of the second working fluid. The second working fluid 254 exits the condenser 256 and enters an expansion valve 258.
The expansion valve 258 can operate as a flow control device within the cooling cycle 250. The expansion valve 258 controls the amount of the second working fluid 254 flowing from the condenser 256 to an evaporator 260. The high-pressure liquid second working fluid 254 exiting the condenser 256 enters the expansion valve 258, which allows a portion of the second working fluid 254 to enter the evaporator 260. The expansion valve 258 allows a pressure drop in the second working fluid 254, thus expanding to a lower pressure prior to entering the evaporator 260.
The expansion valve 258 can have a temperature-sensing bulb filled with a gas similar to the second working fluid 254. The expansion valve 258 opens as the temperature on the bulb increases from the second working fluid 254 exiting the condenser 256. The change in temperature creates a change in pressure on a diaphragm and opens the expansion valve 258. The diaphragm can be biased to a closed position by a biasing element, such as a spring or actuator, and the change in pressure on the diaphragm and causes the biasing element to move the expansion valve 258 to an open position.
The evaporator 260 receives the second working fluid 254 from the expansion valve 258 and allows expansion to a phase that includes both liquid and vapor, with more liquid that vapor. The evaporator 260 passes the second working fluid 254 through to absorb heat from a cooling fluid, such as the fluid 104 from the cooling system 100 of
The evaporator can receive the cooling fluid 104, from the system 100 of
The first working fluid 208 and the second working fluid 254 can be hermetically sealed one from the other within the turbo-compressor 204. The first working fluid can be a thermal fluid optimized for use in the power cycle 202. Representative thermal fluids can include refrigerants, hydrocarbons, inorganic fluids, and/or any combination thereof, which can be operate in the subcritical two-phase region or the supercritical region depending on the waste heat temperature and fluid flow rate and the desired trade-off between compactness and COP. Example subcritical fluids can include refrigerants 1-methoxyheptafluoropropane (HFE-7000), methoxy-nonafluorobutane (HFE-7100), or octafluorocyclobutane (RC318), hydrocarbon propane, or inorganic water or ammonia. Example supercritical fluids include refrigerants octafluoropropane (R218) and carbon dioxide, hydrocarbon ethane, and inorganic xenon.
The second working fluid 254 can be a cooling fluid optimized for use in the cooling cycle 250. Representative cooling fluids can include refrigerants, hydrocarbons, inorganic fluids, and/or any combination thereof, which can be operate in the subcritical two-phase region or the supercritical region depending on the waste heat temperature and fluid flow rate and the desired trade-off between compactness and COP. Example subcritical fluids can include refrigerants 1,1-Difluoroethane (R-152a), pentafluoropropane (R-245fa), 1,1,1,2-Tetrafluoroethane (R-134a), hydrocarbon propane, or inorganic water or ammonia. Example supercritical fluids include refrigerants octafluoropropane (R218) and carbon dioxide, hydrocarbon ethane, and inorganic xenon.
While the first working fluid 208 and the second working fluid 254 can be the same fluid, such as water, the ultra-efficient turbo-compressor cooling system 200 can achieve higher COP utilizing different working fluids. Proposed combinations of the first working fluid and second working fluid can include, but are not limited to, HFE-7100/R245fa; HFE-7000/R152a; RC318/R152a, and R218/R152a, respectively listed as first working fluid/second working fluid.
The power cycle 302 operates with the first working fluid 308 receiving waste heat from a waste heat portion 309 of a co-located process within the facility. In one instance, the waste heat portion 309 is from a co-located process that generates exhaust steam in a steam boiler. The exhaust steam from the steam boiler 311 can be relatively hot (e.g., 300 F). This exhaust heat from the steam boiler 311 may be captured in a glycol-heating loop, and then used in an ORC boiler 313 for heating the first working fluid 308. In another instance, the waste heat portion 309 may be from the ammonia cooling cycle 100. For example, a glycol-heating loop 309 may be used to cool the relatively hot ammonia fluid 104 in the cooling cycle 100 after the discharge of the compressor or the warm lubrication oil in the ammonia compressor. By cooling the ammonia fluid 104, the glycol is heated and, thus, provides the opportunity to be used beneficially in providing heat to the first working fluid 308. The glycol-heating loop 309 may be located at the cooling tower 106 stage of the cooling loop 100 shown in
The turbine 306 has a plurality of vanes configured to rotate as the first working fluid 308 expands within the turbine 306. The gaseous first working fluid 308 exiting the waste heat exchanger 315 enters the turbine 306 and expansion of the first working fluid 308 within the turbine 306 generates mechanical power. The turbine 306 has greater than 80% efficiency in generating mechanical power from the expansion of the first working fluid 308. The mechanical power generated can be transferred to the compressor 352 of the turbo-compressor 304. In the instance of the turbo-compressor 304 including a magnetic synchronous coupling, power loss may be reduced between the turbine 306 and the compressor 352 while hermetically sealing the power cycle 302 and the cooling cycle 350.
Upon exiting the turbine 306, the first working fluid 308 enters a condenser 316. The condenser 216 condenses the first working fluid 308 from a vapor to a saturated liquid by rejecting heat to the environment. Instead of rejecting the waste heat to the environment, the waste heat may be utilized to heat water, among other beneficial uses. As seen in
The first working fluid 308 leaves the condenser 316 as a saturated liquid and enters a mechanical pump 318. The mechanical pump 318 re-pressurizes the first working fluid 308 and circulates the working fluid 308 to the heat exchanger 315 utilizing the waste heat from the glycol-heating loop 309. Still referring to
An expansion valve 358 can operate as a flow control device within the cooling cycle 350. The expansion valve 358 controls the amount of the second working fluid 354 flowing from the condenser 356 to an evaporator 360. The high-pressure liquid second working fluid 354 exiting the condenser 356 enters the expansion valve 358, which allows a portion of the second working fluid 354 to enter the evaporator 360. The expansion valve 358 allows a pressure drop in the second working fluid 354, thus expanding to a lower pressure prior to entering the evaporator 360. In the illustrated embodiment, the second working fluid 354 experiences a pressure drop within the expansion valve 358 and a corresponding saturation temperature drop from 27° C. to less than 6.5° C., allowing the second working fluid 354 to exit the expansion valve 358 at a temperature less than 6.5° C.
The evaporator 360 receives the second working fluid 354 from the expansion valve 358 and allows expansion to a combined liquid and vapor state, with more liquid that vapor. The evaporator 360 is configured to absorb heat from a cooling fluid to the second working fluid 354, thereby generating the desired cooling effect by reducing the temperature of the circulating cooling fluid. In the illustrated embodiment, the circulating cooling fluid is the fluid 104 (e.g., ammonia) from the cooling cycle 100 of
In the illustrated embodiment, the evaporator 360 can receive the circulating cooling fluid 104 at a 27° C. and 140 PSIG and discharge the circulating cooling fluid 104 at 6.5° C. The remaining portions of the cooling cycle 100 can be seen in
As previously described, the fluid 104 leaving expansion valve 110 is a mixture of liquid and vapor. The fluid upstream of valve 110 is all liquid, and the percentage of vapor leaving the valve 110 is dependent on the entering liquid temperature. The higher the entering liquid temperature, the higher the fraction of vapor leaving the valve 110. The vapor portion of the fluid 104 does not participate in the refrigeration process and is simply recompressed. Thus, if the liquid temperature is lowered prior to entering the valve 110, less vapor has to be recompressed in the compressor 102. In a certain instance, with a compressor power of 2.95 MW, the vapor fraction exiting valve 110 is 13.6%, as indicated in
The system 300 of
II. Diesel Generators and Turbo-Compression Cooling System
The turbo-compression cooling system 200 of
Disclosed herein is a turbo-compression cooling system (“TCCS”) driven by low temperature waste heat capable of increasing fuel efficiency of diesel generators by at least 10%. The TCCS is capable of operating within the small footprint associated with the environment. To begin, reference is made to
The DG's 500 provide electricity to various types of ships. Conventionally, the DG's are cooled via jacket water, and lubricating oil. These are sources of waste heat from the DG's. Additional sources of waste heat include exhaust gas; however, using the waste heat in the jacket water and lubricating oil avoids the challenges associated with recovering waste heat from the exhaust gas stream. That said, the exhaust gas is a much higher temperature than the other waste heat streams. While the feasibility of recovering the heat from the exhaust gas is more complicated than recovering waste heat from the jacket water and lubricating oil, it is feasible in certain instances given its relatively high temperature. One challenge to the feasibility of using the exhaust gas as a waste heat source is that it undergoes significant temperature variations (425° C. to 25° C.) and there is exposure to acid contaminants. Furthermore, adding additional heat exchangers imposes additional backpressure on the diesel engine, which may negatively impact engine performance. For these reasons, among others, certain TCCS 504 described herein utilize only the jacket water and lubricating oil as a waste heat source, while others include the exhaust gas as a waste heat source. That is, less than all of the available waste heat streams may be utilized in order to provide efficiency gains to the DG 500 while fitting the overall system within the available footprint within the ships.
Finally, adding a heat recovery device to the exhaust stream will significantly increase the footprint of any system because the heat exchanger is not normally present, as the exhaust heat is simply exhausted to the ambient. In contrast, adding additional pressure drop to the coolant lines will not negatively impact engine performance or significantly increase the footprint of the system. To that end, the TCCS 504 described in this section aims to capture this low-grade heat to achieve a ˜10% fuel economy improvement, while having a less than 4% size and weight penalty.
The jacket water inlet temperature is limited to 90° C. and the jacket water flow rate should be maintained such that the change in temperature is never greater than 11° C. On an exemplary ship and on a per engine basis, the compressor power draw is approximately 14% of the total power provided by each engine. Therefore, if this electrical power requirement can be substantially reduced (or potentially eliminated), then the overall system efficiency can be dramatically improved.
As described herein the TCCS 504 may be integrated with the DG 500 to convert the normally wasted low temperature heat (80 to 90° C.) from DGs 500 into cooling to offload electrical requirements from the chillers 514 that provide chilled water. As shown in
As seen in
For the cooling fluid cycle 518, superheated vapor 538 exiting the evaporator 540 is first heated by the cooling cycle condenser discharge in a suction-line heat exchanger (“SLHX”) 542. This pre-cools the outgoing fluid 538 from the condenser 544, which reduces the inlet enthalpy in the evaporator 540 and substantially improves the performance of the cooling cycle 518. The vapor 538 exiting the SLHX 542 is then compressed to the condenser pressure by the compressor 534. By using the SLHX 542, the compressor 534 discharge temperature increases significantly. However, this extra heat can be rejected to the power cycle 516 in the three fluid recuperator 526. Although preheating the compressor inlet can increase the specific work of the compressor 534, the combination of increased power generated by the turbine 532 and reduced mass flow from the SLHX 542 allows the overall system COP to be high. Incorporation of a SLHX 542 also improves the operational stability of the cooling cycle 518 by preventing liquid droplets from entering the compressor 534. Once the cooling fluid 538 is liquefied in the condenser 544 and the SLHX 542, it is expanded to low pressure and passed into the evaporator 540, where the cooling effect is generated.
One of the challenges with low temperature waste heat recovery is that the system must operate very close to the heat source and heat sink temperatures to have a high heat recovery efficiency. As the saturation temperature of the working fluid gets closer to the source/sink temperature, the heat transfer surface area must increase to accommodate for the lower thermal driving potential. Accordingly, compact heat exchanger technology is utilized to keep the footprint of the TCCS small while still being able to operate very near to the source and sink temperatures. In the TCCS 504 shown in
It is noted that the system 504 is
The compressor and turbine isentropic efficiencies of the TCCS are >80%.
The specific speed of the turbine is proportional to the rotational speed (N in units of RPM) and the square root of the volumetric flow rate (V in units of ft3 s−1) and inversely proportional to the enthalpy rise or drop (H in units of ft. lbf 1bm−1) to the 0.75 power. The volumetric flow rate for compressors and turbines are calculated at the inlet and outlet, respectively. Centrifugal compressors and turbines typically operate with efficiencies in excess of 80% at specific speeds below 120 and 70, respectively. Above these values, radial turbomachines are extremely difficult to manufacture due to their small size (i.e., small tip clearances) and high rotational speeds. For system simplicity, it is ideal for the actual rotational speeds for the power and cooling cycle fluids to be the same. For power plant cooling, using the same fluid is a challenge because the temperature lifts for each of these cycles is substantially different, and the ΔH is significantly larger for the turbine if the same fluid is utilized in the compressor.
In contrast, for the proposed liquid chiller system, the temperature lifts for the two cycles are similar, and the same fluid can be used. However, because turbine exit density is larger than the compressor inlet density for the same fluid, this causes the turbine to have a smaller specific speed (Ns) than the compressor, and care must be taken to select a fluid that falls in the 80% efficiency island on the Cordier diagrams for both components. As shown in
One of many features of the TCCS is there are no mechanical-electrical-mechanical losses to power the compressor, which substantially improves system efficiency. Waste heat from fuel combustion is used to produce mechanical power in an organic Rankine cycle (ORC), which is directly used to drive a compressor on a vapor compression machine, completely bypassing the mechanical-electrical-mechanical conversion losses that currently plague the ship cooling systems. For example, a mechanical transmission efficiency >98% has been measured experimentally for the ˜30,000 rpm TCCS shown in
Another feature of the TCCS is heat integration. As shown in
The combination of the TCCS's efficient turbo-compressor, highly effective heat exchangers, non-corrosive working fluid, and small footprint make it ideal for integration with a marine diesel generator set of a large ship. While there are commercially available thermally activated cooling systems, they suffer inherent limitations that make their use impractical compared to the TCCS described herein.
The magnitude of fuel efficiency increase from using waste heat to power the TCCS depends on how the cooling is used on the ship. Three options for utilization of the waste heat are shown in
Still referring to
Mathematical simulations using average engine loading conditions (57% load) for waste heat data without recovering waste heat from the aftercooler air are as follows. With approximately 575 kW of waste heat input at an inlet temperature of 101° C. The ambient seawater temperature has a large impact on the system performance and the temperature changes as the ship travels throughout its operating locations. The performance was determined for each ambient seawater temperature and then provided a relative weight based on the amount of days the ship would operate in that temperature. Since seawater temperature varies throughout the world, typical annual temperature distributions coupled with engine capacity factors can be used to optimize system performance.
For the purposes of these sample calculations, it is assumed a ship in the open ocean experience 29° C. seawater temperature most often, approximately 55 days of the year. Thus, the thermodynamic performance of each option will be weighted most heavily when the ambient seawater temperature is 29° C. The results of the modeling efforts with only utilizing lubricating oil and jacket water heat in the TCCS are summarized below and in the table in
The option in
The option in
The option in
Preliminary sizes of the three systems shown in
III. Industrial Systems and Turbo-Compression Cooling System
Waste heat from industrial processes is significant and represents an opportunity to improve energy efficiency of manufacturing processes. For example, waste heat from ultra-low temperatures from jacket cooling water engines (about 90° C.) and waste heat from steam boilers (about 150° C.) are available and largely untapped recovery streams. Thus, the following systems are generally applicable to a wide variety of industrial processes and/or manufacturing environments. In addition, while the previous sections of this application described specific processes for specific applications (e.g., marine diesel generators), the turbo-compression cooling systems described in those section are equally applicable to more generalized industrial processes and/or manufacturing environments. In addition, while each system is discreetly described in this application, it is to be understood that features from one system may be applied to the other systems described herein without limitation. Each system embodiment is merely illustrative of an example embodiment and is not intended to be limited to its specific illustration.
To begin, reference is made to
The waste heat boiler 614 receives ultra-low temperature waste heat from an industrial process and thus heats the power cycle fluid 612 prior to the fluid 612 entering the turbine 616. As stated previously, the waste heat boiler 614 may receive heat from engine jacket coolant and/or steam boilers, among other equipment in the facility.
Turning to the cooling cycle 608, the fluid 622 leaving the economizer 604 enters the condenser 626, which may be a water-cooled condenser. The fluid 622 may then enter a first passage of a suction-line heat exchanger (“SLHX”) 628. The fluid 622 in the first passage pre-heats the fluid 622 passing through a second passage of the SLHX 628 prior to that fluid 622 entering the compressor 624. Once the cooling fluid 622 is liquefied in the condenser 626 and the SLHX 628, it is expanded to low pressure and passed into an evaporator 630, where the cooling effect is generated. The evaporator 630 may be used in a cooling process within the facility such as, for example, chilling water or another substance. Superheated vapor 622 exiting the evaporator 630 passes through the second passageway and is heated by the SLHX 542. This also pre-cools the fluid 622 in the first passageway of the SLHX 628, which reduces the inlet enthalpy in the evaporator 630 and substantially improves the performance of the cooling cycle 608. Incorporation of a SLHX 628 can also improve operational stability of the cooling cycle 608 by preventing liquid droplets from entering the compressor 624.
On the cooling cycle 706 side, a second compressor 720 may be included in the system 700, pre-compressing the fluid 708 right before it is received by the first compressor 714. The second compressor 720 may be electrically powered, whereas the first compressor 714 is powered by the operation of the power cycle 702. More particularly, the power cycle 702 is powered by a waste heat stream 718 from a co-located industrial process. The waste heat stream 718 may be an ultra-low temperature stream of waste heat including engine jacket cooling water and lubricating oil (e.g., generator). The heated power cycle fluid 704 is used to power the turbine 712 of the turbo-compressor 710, which powers the first compressor 714. The second compressor 720 is powered with electricity to increase the efficiency of the system 700. In this way, the cooling cycle 706 operates two compressors 714, 720, which eliminates surging of the compressor 714 while still cooling primary chilled water 724 to the same inlet and outlet temperature when the availability of waste heat 714 decreases.
Referring back to
The recuperator 730 additionally facilitates pre-heating of the power cycle fluid 704 by the discharge line coming off the turbine 712 passing through the recuperator 730 prior to entering the condenser 736. After passing through the condenser 736, the fluid 704 is pumped via a pump 738 back through the recuperator 730 and to the waste heat boiler 732 before entering the turbine 712. It is noted that the condensers 736, 734 in this embodiment, and others, may be combined into a single condenser. Moreover, in embodiments with the same cycle fluid 704, 708, the fluids may be mixed at certain points throughout the system 700.
With the system 700 shown in
It is noted that elements of the systems shown and described with references to
The graph of
It is believed the exemplary embodiment and its advantages will be understood from the foregoing description, and it will be apparent that various changes may be made thereto without departing from the spirit and scope of the disclosure or sacrificing all of its advantages, the examples hereinbefore described merely being preferred or exemplary embodiments of the disclosure.
This application claims the benefit of U.S. Provisional Application No. 62/941,570, filed Nov. 27, 2019, which is hereby incorporated by reference in its entirety into the present application.
This invention was made with government support under DE-AR0000574 and DE-EE0008325 awarded by the Department of Energy. The government has certain rights in the invention.
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