US3788066A - Refrigerated intake brayton cycle system - Google Patents

Refrigerated intake brayton cycle system Download PDF

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US3788066A
US3788066A US3788066DA US3788066A US 3788066 A US3788066 A US 3788066A US 3788066D A US3788066D A US 3788066DA US 3788066 A US3788066 A US 3788066A
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air
refrigeration
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brayton cycle
water
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W Nebgen
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BRAYTON CYCLE IMPROVEMENT ASS
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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02CGAS-TURBINE PLANTS; AIR INTAKES FOR JET-PROPULSION PLANTS; CONTROLLING FUEL SUPPLY IN AIR-BREATHING JET-PROPULSION PLANTS
    • F02C3/00Gas-turbine plants characterised by the use of combustion products as the working fluid
    • F02C3/20Gas-turbine plants characterised by the use of combustion products as the working fluid using a special fuel, oxidant, or dilution fluid to generate the combustion products
    • F02C3/26Gas-turbine plants characterised by the use of combustion products as the working fluid using a special fuel, oxidant, or dilution fluid to generate the combustion products the fuel or oxidant being solid or pulverulent, e.g. in slurry or suspension
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01KSTEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K17/00Using steam or condensate extracted or exhausted from steam engine plant
    • F01K17/06Returning energy of steam, in exchanged form, to process, e.g. use of exhaust steam for drying solid fuel or plant
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01KSTEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K21/00Steam engine plants not otherwise provided for
    • F01K21/04Steam engine plants not otherwise provided for using mixtures of steam and gas; Plants generating or heating steam by bringing water or steam into direct contact with hot gas
    • F01K21/047Steam engine plants not otherwise provided for using mixtures of steam and gas; Plants generating or heating steam by bringing water or steam into direct contact with hot gas having at least one combustion gas turbine
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02CGAS-TURBINE PLANTS; AIR INTAKES FOR JET-PROPULSION PLANTS; CONTROLLING FUEL SUPPLY IN AIR-BREATHING JET-PROPULSION PLANTS
    • F02C7/00Features, components parts, details or accessories, not provided for in, or of interest apart form groups F02C1/00 - F02C6/00; Air intakes for jet-propulsion plants
    • F02C7/12Cooling of plants
    • F02C7/14Cooling of plants of fluids in the plant, e.g. lubricant or fuel
    • F02C7/141Cooling of plants of fluids in the plant, e.g. lubricant or fuel of working fluid
    • F02C7/143Cooling of plants of fluids in the plant, e.g. lubricant or fuel of working fluid before or between the compressor stages
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B11/00Compression machines, plant, or systems, using turbines, e.g. gas turbines
    • F25B11/02Compression machines, plant, or systems, using turbines, e.g. gas turbines as expanders
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02GHOT GAS OR COMBUSTION-PRODUCT POSITIVE-DISPLACEMENT ENGINE PLANTS; USE OF WASTE HEAT OF COMBUSTION ENGINES; NOT OTHERWISE PROVIDED FOR
    • F02G2250/00Special cycles or special engines
    • F02G2250/03Brayton cycles
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B2400/00General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/14Power generation using energy from the expansion of the refrigerant
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B2400/00General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/14Power generation using energy from the expansion of the refrigerant
    • F25B2400/141Power generation using energy from the expansion of the refrigerant the extracted power is not recycled back in the refrigerant circuit

Abstract

In an open Brayton cycle, suction air is cooled by refrigeration to a predetermined temperature preferably to at least 50*F. below ambient air temperature. In the case of a conventional refrigeration system with refrigerant evaporation and condensation utilizes ambient air for condensation, the refrigeration is controlled to maintain a fixed temperature differential between the temperature at which the refrigerant evaporates and at which it condenses. If the temperature of the suction air is below the freezing point of water an aqueous solution of a water vapor pressure depressant such as glycol may be used under conditions which avoid, or control, freezing. This aqueous solution may be sprayed on cooling coils in indirect heat transfer or it may be cooled by circulation in indirect heat transfer and then used in direct contact to cool and condense water vapor from the suction air. The condensed water is vaporized, preferably with waste heat and injected into the compressed air stream. The refrigeration system may be a conventional one driven by the shaft or from the power network or it may be one which is driven by a system utilizing waste heat of the Brayton cycle exhaust gases. In a conventional system, multistage refrigeration is preferred. In the use of a solution of glycol the composition is maintained by distillation of water under pressure with utilization of the steam in the air stream between the Brayton cycle compressor and expander to improve output and thermal efficiency. The advantages of suction air refrigeration are especially marked when in combination with regeneration in the Brayton cycle and the advantages of regeneration are greater when in combination with suction air refrigeration. Still greater advantages result from a combination of suction air refrigeraton with both regeneration and gasification of fuels which contain sulfur and particulates. A portion of the refrigeration capacity for suction air cooling may be used for the cooling of electric generators and transformers. Suction refrigeration results in similar advantages when used with a closed Brayton cycle, wherein the working fluid is a dry gas, such as helium, or argon, or the like.

Description

United States Patent [191 Nebgen [54] REFRIGERATED INTAKE BRAYTON CYCLE SYSTEM [58] Field of Search 62/93, 208, 209, 210, 211,

[56] References Cited UNITED STATES PATENTS 1,879,685 9/1932 Jaczko 60/39.67 2,548,508 4/1951 Wolfner 60/3967 X 2,678,531 5/1954 Miller 60/3955 3,621,656 11/1971 Pacault et al. 60/3967 X 3,649,469 3/1972 McBeth 60/3955 X- 2,216,690 10/1940 Madden..... 62/208 X 2,734,346 2/1956 Dickieson, Jr. 62/217 X 2,888,809 6/1959 Rachfal 62/208 X 3,250,084 5/1966 Anderson... 62/208 X' 3,260,064 7/1966 Newton....' 62/217 X 2,339,185 l/l944 62/238 X Nettel Primary ExaminerWilliam FIODea Assistant Examiner-Peter D. Ferguson Attorney, Agent, or Firm-Robert Ames Norton et al.

CONDENSE D WATER 3 Jan. 29, 1974 maintain a fixed temperature differential between the temperature at which the refrigerant evaporates and at which it condenses. If the temperature of the suction air is below the freezing point of water an aqueous solution of a water vapor pressure depressant such as glycol may be used under conditions which avoid, or control, freezing. This aqueous solution may be sprayed on cooling coils in indirect heat transfer or it may be cooled by circulation in indirect heat transfer and then used in direct contact to cool and condense water vapor from the suction air. The condensed water is vaporized, preferably with waste heat and injected into the compressed air stream.

The refrigeration system may be a conventional one driven by the shaft or from the power network or it may be one which is driven by a system utilizing waste heat of the Brayton cycle exhaust gases. In a conventional system, multistage refrigeration is preferred.

The advantages of suction air refrigeration are especially marked when in combination with regeneration in the Brayton cycle and the advantages of regeneration are greater when in combination with suction air refrigeration.

Still greater advantages result from a combination of suction air refrigeraton with both regeneration and gasification of fuels which contain sulfur and particulates.

A portion of the refrigeration capacity for suction air cooling may be used for the cooling of electric generators and transformers.

Suction refrigeration. results in similar advantages when used with a closed Brayton cycle, wherein the working fluid is a dry gas, such as helium, or argon, or the like.

10 Claims, 3 Drawing Figures BRAYTON CYCLE AIR COMPRESSOR COOLED AIR EXPANDER BRAYTON CYCLE R EC UP E RATOR COMBUSTION CHAMBER HOlVBOdVAH CONTROLLER AMBIENT AlR o REFRIGERANT COMPRESSOR 67 AMBIENT AIR PATENTEDJAII 29 I974 BRAYTON CYCLE SHEEI 1 OF 3 BRAYTON CYCLE AIR COMPRESSOR EXPANDER c 1 GENERATOR Ep TT AT T REc A I R TOO COOLER T RECUPERATOR m II /I s E' FUEL REFRIGERANT L HEATER T INTAKE AIR NH I EXHAUST I 7- C C Is I II 29 28 27 I5 I4 I3 I II 33 23 7 I2 HOT 34 I STORAGE REFRIG REFRIG TANK COMPRESSOR EXPANDER REFRIG V CONDENSER f 3 4 HOT FLASH TANK AMBIENT STORAGE TANK INVENTOR WILLIAM H. NEBGEN ATTORNEY PATENTEDJIIII 29 I974 7 SHEET 2 OF 3 I20 PM; i sTEAM As REOO. FOR PROCESS COMPRESSED AIR Ioo PSIG 1 43 PRODUCT 22 5OF 700F I ,sTEAM TURBINE 4| PARTIAL o FUEL OXIDATION i F BOILER GENERATOR REAOTOR 1 WATER L STEAM 39 CONDENSER 50 I RECYCLE BF PUMP BLOWER Fw IsoFi PARTIcuLATE HEATER SCRUBBER H25 REMOVAL EXPANDER COM/PIZESSOR 58 52 F g; 5 GENERATOR o NR 800 cOMBusTOR o REFRIGERATED 900 5 A ATMOSPHERIC 54 AIR r r RECUPERATOR REOOPERATOR PURIFIED FUEL GAS EXHAUST 60 J imolt ISOOFt INVENTOR WILLIAM HRNEBGEIN FIG. 2

ATTORN Y -1 REFRIGERATED INTAKEBRAYTON CYCLE SYSTEM 'RELATEDYAPPLICATION This is a continuation-in-part of my co-pending application, Ser. No. 34,717, filed May 5, 1970, now U.S. Pat. No. 3,668,884 dated June 13, 1972.

BACKGROUND OF THE INVENTION In a gas turbine the power output depends on the suction air temperature and increases as the temperature is lowered, other parameters remaining the same. This suction temperature is normally that of the ambient air and fluctuates daily, seasonally and with atmospheric conditions.

When the air entering the suction of a given engine is cooled by refrigeration the compression ratio and mass flo'w increase and the expansion ratio and mass flow also increasefso thatthe net shaft output is increased accordingly. This increase in power output is considerably more than the power that is needed to effect the refrigeration. For the combined system of the Brayton cycle and refrigeration cycle the thermal efficiency is not significantly different from that of the Brayton cycle alone. The additional capital expense for refrigeration is often less than the value of the increase in power output that results from refrigerating the suction air. g

In an open Brayton cycle engine, air enters the engine at atmospheric pressure, is compressed, is heated by being burned with fuel and then is expanded back to atmospheric pressure. The net work output of the engine is the relatively small difference between two quite large numbers, i.e., it is the difference in the total work' produced by its expander and the work con- 'sumed by its air compressor. In this discussion, the

open Brayton cycle expander for convenience is referred to as an -air expander, although the working fluid actually contains the products of combustion of the fuel. The work produced by the air expander of a 5.4 ratiosimple Brayton cycle engine is about 2.77 times the net work output of the engine, and when the compressor takes suction at ambient temperature (for example 100F.) the work consumed by the air compressor is about 1.77 times the network output. If the ambient temperature air is refrigerated before it enters the compressor, the work output of this Brayton cycle engine increases because the compression ratio increases, and the expansion ratio increases accordingly; the work produced by the air expander therefore increases; and also because the mass flow of air through the engine increases, due to the greater density of the cold air. The work which is required to refrigerate the' inlet air must, of course, be deducted from the work which is produced by the Brayton cycle engine, but even when an inefficient single stage refrigeration system is used, the refrigerated suction engine delivers more usable shaft work than does the same engine if it takes suction at ambient temperature.

Similar advantages result from refrigerating thev suction of a closed Brayton cycle engine, wherein the working fluid may be dry, and is heated indirectly in an external heat exchanger.

Refrigeration of suction air presents certain problems which arise fromthe condensation of moisture as the temperature of the air is reduced below the dew point.

In the case of a large power facility the quantity of moisture is very large. For example, a typical facility to generate 200 megawatts, and receiving ambient air at F. and 50 percent relative humidity, requires removal of 325,000 gallons daily of water when the dew point is reduced to 32F.

Most of this is recoverable as liquid water by cooling the suction air to a temperature which is in the vicinity of the freezing point. Normally this requires refrigeration, for example, by indirect heat exchange with an evaporating refrigerant. Water condenses on the heat exchange surfaces and the runoff is collected and recovered. If the air were to be cooled by refrigeration substantially below the freezing point the vapor would condense to produce ice, or rime, on the cooling surface. This accumulation in a short time would block the passages for air flow in indirect heat transfer apparatus for air cooling.

In the prior art of cooling air, for example the refrigerated storage of food, icing is prevented or reduced by applying to the heat'exchange surface a liquid which SUMMARY OF THE INVENTION One aspect of the present invention concerns an improved method of heat transfer for cooling and dehumidification of suction air to an open Brayton cycle. In one embodiment of this method the suction air is caused to pass over a heat transfer surface which is preferably wetted by an aqueous solution, e.g. of ethylene glycol, methanol, etc. The composition of the aqueous solution which is applied to the surface is controlled to avoid solidification on the heat transfer surface at the temperature of the refrigerant or cold fluid.

It is, of course, clear that the aqueous solution absorbs the water vapor from the suction air and becomes diluted thereby. In accordance with this invention the composition and quantity of the solution applied to the heat transfer surface is controlled, in combination with the quantity of water condensed, so that the composition of the solution when it is diluted with condensate is everywhere on the heat transfer surface a composition corresponding to a freezing point which is sufficiently below the temperature of the refrigerant or coolant to prevent ice deposition.

Usually itis advantageous to cool the air in stages to the desired suction temperature. The minimum concentration of the aqueous solution in contact with the Ultimately the final spent solution is regenerated by I a distillation separation obtaining a more concentrated aqueous solution to be recycled for application on the heat transfer surface. When, as for example in the case of a glycol, the aqueous solute is less volatile than water the latter is removed in the overhead vapor of the distillation whereas, for example in the case of methanol,

the solute, being the more volatile component, is recovered in the overhead fraction, leaving the water in the bottoms. From this overhead vapor the methanol is condensed and recycled.

The choice of an aqueous solute in accordance with this invention depends on the temperature range of the air cooling stage. Ethylene glycol is a preferred solute, at temperatures above about 40F. Below this temperature the viscosity of the aqueous solutions of glycol is high. At lower temperatures methanol provides the desired freezing point depression without excessive viscosity or too high vapor pressure whereas at higher temperatures methanol is too volatile.

In a second embodiment of this invention there is direct transfer of heat from the air to the aqueous solution at each stage of cooling, together with indirect heat transfer from the aqueous solution to a refrigerant or coolant. The aqueous solution is recycled between these two heat exchange operations and it serves thereby as a medium for heat exchange as well as for absorption of condensed water vapor from the air. The concentration of the aqueous solution in each cooling stage is controlled by the withdrawal of dilute solution and the return of a more concentrated solution to replace the dilute solution whichis withdrawn. The difference in the water content of the dilute solution and the concentrated solution represents the water vapor that has been condensed from the air.

In this embodiment, as in the first embodiment, the composition of the aqueous solution is controlled so that at the temperature of the refrigerant or coolant, freezing does not occur on the heat transfer surface which in this embodiment separates the refrigerant, or coolant, and the aqueous solution.

Contact between the aqueous solution and the suction air is by means of a packed bed or other apparatus and the scope of the invention is not limited to any particular form of apparatus. The transfer from the aqueous solution to the refrigerant or coolant is preferably in a shell and tube heat exchanger but again the invention is not limited thereto.

At any stage of air cooling which is above the normal freezing point of water'there is no need for an aqueous solution to control the freezing on the heat transfer surface in either embodiment of this invention and water could be recirculated in the second embodiment. However, the aqueous solution of either embodiment dehumidifies as well as cools the air. For this reason it is advantageous to utilize an aqueous solution even in the higher temperature stages of cooling above 32F. since this tends-to remove, at a given temperature, a larger amount of water vapor from the air, and thus reduces the work of refrigeration. Another advantage of using an aqueous solution of a freezing point depressant even where cooling would not result in ice deposition is that it is not necessary to change the composition of the recirculated coolant as temperatures change and this simplifies equipment and operation.

While a freezing point depressant is preferred, the present invention in its broader aspects is not absolutely limited to such a material for the prevention of formation of solid ice on heat exchange surfaces. It is possible to use a material which is relatively immiscible with water, for example, a liquid hydrocarbon. In this case the freezing point of condensed water is not actually depressed but the flow of the substantially nonsolvent liquid keeps ice crystals very small and in effect keeps them in a dispersed form so that they do not deposit as solids on the refrigerating equipment. Separation of water from a non-solvent can be effected simply and economically by raising the temperature above the freezing point of water, forming liquid water, which can be separated by decantation or other conventional methods of separating water from non-solvent liquids. It will be noted that regardless of how ice formation is controlled, relatively pure water is produced.

Another aspect of this invention pertains to the utilization of condensate water in an open Brayton cycle for improvement of thermal efficiency and of power output capacity. One method to use this condensed water is to spray it into the air compressor of the Brayton cycle and another method of utilization is to inject the steam resulting from the vaporization of the condensate into the stream of air from the Brayton cycle compressor. It should be noted that increase of thermal efficiency and power output capacity do not necessarily involve the same economic considerations. Thermal efficiency increases are largely factors which lower fuel cost. However, for certain uses, such as Brayton cycle installations for power peaking in electric generating plants, increases in power output may be more valuable than savings in fuel cost. As has been pointed out above, and will appear below, not all of the features of the present invention increase both thermal efficiency and power output capacity. In the case of the use of condensate water, there is the fortunate situation that both factors are improved. It should be noted that the present invention in the aspect just set out need not be limited to using all of the condensed water in the Brayton cycle and that the invention, therefore, does include combinations of features in which only part of the condensate water is used in the Brayton cycle. However, as the amount of condensate water is normally less than that which can be effectively used in the open Brayton cycle, it is usually preferable to use all of the condensate water.

If the water is utilized as steam to increase the volume of gas to the expander, the. steam must be generated at a pressure equal to, or slightly higher than, the Brayton cycle compressor discharge and it should be admitted at, or prior to, the combustor. The heat that is needed to generate the steam preferably is obtained by waste heat recovery from the stack gases. Condensate which is obtained as liquid water, either from condensation above the temperature of ice formation, or by recovery from an aqueous solution of a solute which has volatility differing from that of water, is evaporated under a pressure which is at least as high as that of the Brayton cycle compressor discharge. When the condensate water is in the form of a solution of glycol or othersolute which is less volatile than water the water is distilled at the pressure required for injection and the vapor is rectified to the extent necessary to minimize the loss of solute. The rectified water vapor represents the steam which is then suitable for injection, the distillation and rectification heat preferably being obtained from the Brayton cycle exhaust, though other sources of waste heat may be used.

This invention is not limited to any particular refrigeration system, but a multistage refrigeration system in many cases is preferable since it is more efficient than a single stage system because all of the heat withdrawal is not at the lowest temperature.

Still another aspect of this invention relates to the improvement which results from the combination of refrigeration of suction air at least 50F. below ambient with recuperation of the waste heat from the turbine exhaust in a regenerator. Recuperation or regeneration to improve the energy efficiency of a simple Brayton cycle by utilizing some of the waste heat to reduce the amount of fuel required to raise the gases to the turbine inlet temperature is not unknown, and it is an advantage of the present invention that recuperation can be used, and used even more effectively, than in the prior art.

The quantity of heat recuperated is limited by the temperature difference between the turbine exhaust and the compressed air which absorbs the heat. When as in accordance with this invention, the suction temperature is reduced by refrigeration of the air, the temperature of the compressed air is lowered. This increases the capacity of the compressed air to absorb waste heat from the exhaust gases. There is an energy saving of one Btu for every Btu of transfer in the regenerator. Consequently the combination of heat regeneration and refrigeration to reduce the compressor suction temperature is an important advantage of this invention.

Still another aspect of this invention relates to keeping a substantially constant differential between the temperature of the suction air and that of the ambient air, regardless of the ambient air temperature. This permits optimum utilization of the refrigeration system to improve the power output of theBrayton cycle and to maintain the optimum improvement regardless of weather conditions.

In accordance with this last aspect of the invention, I employ the ambient air as the heat sink of the refrigeration system, i.e., as the coolant of the refrigerant condenser. The vapor pressure in the condenser varies with the temperature of the coolant air, and this represents a corresponding variation of the discharge pressure of the refrigeration compressors. Since it is desirable that the suction air temperature be a fixed amount below the ambient air temperature, the pressure in the evaporator also varies, the pressure being higher with a higher ambient air temperature. The refrigerant com-- pressor is a constant volume device, but the density of the refrigerant vapor varies inversely with its pressure so that, without suitable control, the mass flow of refrigerant is similarly variable. As a consequence, the

compressor of a system which is designed to cool the suction air by, say, 50F., at an ambient temperature of 95F, will not have the capacity tocool the air by the same amount when the ambient temperature is 40F. Accordingly, in the system of this invention, I provide a means for throttling the vapor from the evaporator so that at the higher temperatures the flow of vapor is reduced, while at lower ambient temperatures it is in-.

creased to compensate for the inverse tendency which is a consequence of the changes of vapor density.

One method of throttle control is by adjustable guide vanes in the inlet to the compressor. The angle of the vanes and the space between them is adjusted manually or automatically to control the flow of the refrigerant vapor and, thereby, the rate of evaporation, so that there is maintained a substantially constant difference betweenthe temperature of the Brayton cycle suction air and that of the ambient air. The signal output of two temperature sensors, one in the ambient air and the other in the compressor suction, canbe used to operate a servo system which positions the inlet guide vanes.

In a multistage system of refrigeration in which the mass flow'of refrigerant in each compression stage is dependent on the other stages, the preferred embodiment is one which provides throttle control at each stage. For example, when the range of ambient variation is 50F at the maximum ambient temperature the compressor should be throttled to about one-third of its capacity and should be wide open at the lowest ambient temperature.

When it is desired to refrigerate the suction air to temperatures as low as -40F. or less it becomes difficult to find non-volatile substances which, in aqueous solution, are sufficiently fluid at .these low temperatures. Accordingly, it is an object of this invention to remove moisture from the air by means of an aqueous solution at the next higher temperature stage of refrigeration so that the air which enters the lowest temperature stage of refrigeration has a dew point which is so low that the final stage may be by means of a heat transfer surface which remains dry, i.e., free of ice condensation without the application of an aqueous solution. I

This is achieved, in accordance with this invention, by control of the concentration and temperature of the aqueous solution in contact with the air in the next to final stage of refrigeration, in combination with the final stage temperature. The. advantage gained from this method of staging of the air cooling is that the aqueous solution does not have to be used at the lowest temperature of the air, at which the solution viscosity may be excessively high. For example, a 60 percent glycol solution is used at -32F. to cool, and remove moisture from, air whichmay then be further cooled dry to 45F. At 32F. the viscosity of the aqueous solution of glycol is centipoises, which is not too heavy for process use, whereas at'the lower temperature of -,45F. the viscosity is 500 centipoises, which is excessive. The control of these temperatures in combination with the aqueous solution composition thus enables the cooling of the air to a temperature which is not otherwise practical.

Still another aspect of this invention relates to the utilization of dirty fuels as the source of energy isan open Brayton cycle. In a-prefer'red embodiment suction air to the compressor is refrigerated to at least 50F. below ambient. The compressed air is divided, into two streams, one stream for the partial combustion and gasification of the dirty fuel, and the other for the combustion in the Brayton cycle engine of the gasified fuel after it has been cleaned.

Gasification and combustion occur at a temperature above 2,200F. and are carried out in apparatus which is Well known and which may be suitable for either coal or oil. One known method and apparatus is described in the Khristianovich et al. US Pat. No. 3,287,902, Nov. 29, 1966, to which reference is hereby made. The patent, while making general reference to the use of a gas turbine, is primarily directed to the combination of the cleaning of the dirty fuel with a steam turbine. There is, of course, no suggestion of the other features of the present invention, namely the combination with refrigerated suction, compressors, etc. Prior to combustion the compressed air is heated by regeneration utilizing the waste heat of the Brayton cycle turbine exhaust. When air is used in the gasification of fuel, the ratio of air to fuel must be high enough to maintain the reaction temperature. Subject to this limitation, the ratio should be minimized to retain the maximum heat of combustion in the products. The heat which is absorbed by the combustion air in regeneration contributes to a reduction of the air requirement and a corresponding increase of heat retention.

The products of partial combustion are cooled in a pressurized waste heat boiler and boiler feed water heater and are scrubbed and cleaned to remove particulates and sulfur compounds or, in the case of some oils, nickel and vanadium compounds. The cleaned gases then are heated in a section of the regenerator, flowing in parallel with the air from the compressor. The heated cleaned gases are burned with preheated air in the combustors, and the hot products of combustion then undergo expansion in the turbine in the case of an open Brayton cycle or the heat exchanger heating a working fluid in the case of a closed Brayton cycle.

The system of this invention provides unique advantages in efficiency. By refrigeration of the suction air to the compressor, the compressed air to the regenerator is cooler and has a much greater capacity for absorption of waste heat than does the equivalently compressed ambient temperature air. The cooled products of fuel gasification offer additional capacity for heat absorption, and as a consequence, the exhaust gases from the regenerator are considerably cooler than in a system that does not regeneratively heat the gasification products. I

The purified fuel gases, when used in a Brayton cycle, suffer no disadvantage by having their caloric content so greatly reduced because of dilution by the nitrogen of the partial oxidation air. These fuel gases, when preheated by the waste heat available from the regenerator of the Brayton cycle, have a flame temperature which is considerably higher than that which the Brayton cycle turbine can endure, and excess airis still needed for quenching, although not to the same extent as when higher heat content fuels are used. All of the allowable temperature of the purified low heat content fuel gas is, therefore, effectively used in the Brayton cycle, whereas its relatively low flame temperature makes it uneconomical to use for many other purposes, such as for steam generation. In other words, the refrigerated suction and regenerated air and fuel gas features of the present invention results in being able to use effectively in a Brayton cycle fuel gas of low heat content. In the present application the broad combination of production of clean, lower heat content fuel gas from dirty fuel with a Brayton cycle generally is not covered since this broad application is described in my co-pending application Ser. No. 44,673. It is, however, an advantage that the other features of the present invention can be used with the partial oxidation of dirty fuel, which in a number of cases permits further economies in fuel cost.

faced with a number of problems: One is that of peak power, which may be required for only a few hours a day and/or a few days a year. Brayton cycle systems can start up quickly and are ideal for peaking power purposes. Another factor is the increasing stringency of regulation for environmental pollution. Stack gas cleanup processes are currently being developed, but they all are expensive, so at present the most satisfactory method of pollution control is to use clean fuels. An open Brayton cycle requires quite clean fuel in any event because some pollutants are intensely corrosive to most metals and other materials in a gas turbine or heat exchangers for the working fluid of a closed Brayton cycle at the temperatures at which the turbine or heat exchanger operates, whereas the same pollutants frequently do not present such severe corrosion problems at the much lower temperatures of the surfaces of steam boilers. Therefore, Brayton cycles, particularly open Brayton cycles, have to use relatively clean fuel regardless of environmental pollution requirements. This has presented an interesting economic factor in the use of Brayton cycle engines in power plants. For environmental reasons the fuels used in any generating plant must be clean, and since Brayton cycle plants, such as open Brayton cycle plants, are usually much cheaper than steam plants, it is common practice in many large central plants to use Brayton cycle plants, which were originally intended for peak power production, for a much longer portion of the year. The increases in Brayton cycle thermal efficiencies and power outputs made possible by the present invention therefore assume great economic importance in this large and growing field.

Both thermal efficiency and maximum power output are important. There will be described below a preferred embodiment in which the degree of cooling and the compression ratios are optimized for maximum efficiency. In a more specific aspect of the present invention such optimized systems or close approaches thereto, which will be described below,'are included. However, in broader aspects the present invention includes the essentials of cooling by means of multistage refrigeration systems. It is an advantage of the present invention that it can'be used in various degrees of sophistication and optimization depending on the best economic compromise for a particular use or a particular installation.

Maximum refrigeration efficiency is obtainable by a refrigerating system which forms the subject matter of the elected invention of the parent application, Ser. No. 34,717, above referred to. This system is there referred to as the Treadwell System, and this nomenclature will be used in the same sense in the present application. This most highly effective refrigeration system in combination with the other features of the pres- I ent invention is included in a more specific aspect of the present invention, though the refrigeration system as such is, of course, not claimed herein. The combination of the principal features of the present invention with the Treadwell system of refrigeration is in no sense an aggregative use of two isolated features or subcombinations of features.

When the present invention is used for electric power generation, the largest single field at present, as has been set out above, a further refinement is possible and is of practical importance in some cases. Because of the large quantities of incoming air, a relatively large refrigeration system is needed. A unit of refrigeration capacity in such large systems is much more economical carbon. These fluids can be cooled by a portion of the capacity of the large refrigeration system which is used for cooling the incoming air in the Brayton cycle. This utilization of a very small part of the large refrigerating capacity permits the economical sub-ambient cooling and the consequent increase in capacity of generators,

transformers, etc. It is also necessary to cool lubricating oil in Brayton cycle turbines, and the same considerations apply'as this can be done more economically by the relatively very large refrigeration system necessary in the present invention.

In thedescription of the preferred embodiment the description of the particular Treadwell system of refrigeration will be repeated as in the parent case, but, as has been pointed out above, this is only one typical refrigeration system albeit one capable of giving optimum results, and the invention is not limited in its broader scope to the particular Treadwell system but includes any refrigerating system having the limitations of cooling and prevention of ice or other solids formation which has been discussed in detail above. I

Most of the preceding discussion has been directed Particularly to a refrigerated suction regenerated open Brayton cycle engine. Very similar improvements in output and efficiency result when refrigeration and regeneration are applied to a closed Brayton cycle engine, in which the cooling fluid, which may be a dry gas such as helium, argon, etc., is heated indirectly in an external heat exchanger. Such a closed Brayton cycle is sometimes used with atomic power generators.

Because of the greater flexibility in plant arrangement, significant economic advantage is obtained when an auxiliary fluid, such as Dowtherm, is used to transfer regeneratedheat from the expander exhaust to the compressor discharge.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 shows in purely diagrammatic form the combination of cooling the suction of an air compressor for a regenerated open Brayton cycle, and shows as the refrigerating system the Treadwell system; I

FIG. 2 shows in purely diagrammatic form a system for the use of dirty fuel ina Brayton cycle, and

FIG. 3 is a diagrammatic showing of a conventional refrigeration system refrigerating the intake to the air compressor of a Brayton Cycle System.

DESCRIPTION OF THE PREFERRED EMBODIMENTS FIG. 3 illustrates diagrammatically a simple, conventional refrigerating system-with refrigerant evaporation and condensation, the latter being by heat exchange with ambient air. Ambient air passes through the refrigerant evaporator 61 where it is cooled to a predetermined temperature, such as at least 50F. below the temperature of the ambient air, in the presence of a freezing point or vapor pressure depressant in the form of an aqueous solution when a predetermined temperature is sufficiently low to present the possibility of water from the ambient air condensing out and freezing. Water condensed out is'passed through a conduit to a point in the Brayton cycle system between air compression and turbo expansion. This will be described in more detail below. The air passing through the evaporator 61, and cooled therein, is shown leaving through a conduit at the top of the evaporator. As indicated, this cool air goes to the inlet of a Brayton cycle air compressor, which is shown diagrammatically as the compressor is a conventional piece of apparatus and the exact mechanical construction of a particular compres-' sor forms no part of the-present invention, which includes the use of any well known Brayton cycle air' compressor or compressor system.

The refrigerant vapors leave through conduit 62, passing through a throttling valve 75 and thence through the conduit 63 to a refrigerant compressor 64 driven by a motor 65. The compressed refrigerant vapors pass through conduit 66 into the top of the condenser 67 where they are cooled by heat exchange by ambient air as indicated. This results in condensation of the refrigerant vapors to liquid refrigerant which passes through pipe 68 and expansion valve 76 into the refrigerant evaporator 61.

There are two temperature sensors, one 69 for the ambient air entering the evaporator 61 and the temperature therein, which is sensed by another sensor 70. These sensors connect to inputs 71 and 72, which may be wires, to a controller 73 which converts them to a differential signal 74 which operates a servo mechanism 76 driving the throttling valve 75. The controller 73, of conventional design, is set for the predetermined temperature between sensors 69 and which sense the temperature of refrigerant condensation and evaporation respectively. The temperature of the ambient air is, of course, substantially that of the condensed refrigerant in a condenser 67. As is common in differential controllers, controller 73 will put out a signal 74 only when the differential between the temperature sensed by sensor 69 and sensor 70 depart from the predetermined value for which the controller 73 is set. As is customary in such control circuits, the signal 74 is in a different phase depending on whether the temperature differential between sensors 69 and 70 is greater or less than the predetermined value set in the controller. This signal operates the servo mechanism 76 to vary the setting of the valve 75 and hence the throttling of the refrigerant vapor. As soon as the throttling has brought the temperature differential between sensor 69 and 70 to the predetermined value, the signal 74 ceases and the throttling therefore maintains the predetermined temperature differential regardless of the temperature of the incoming ambient air.

In the Summary of the Invention a throttling method with adjustable guide vanes is described. This, of course, performs precisely the same function as a simple throttling valve 75. The latter, however, is simpler to illustrate in a purely diagrammatic drawing such as FIG. 3. The invention is, of course not limited to any particular special design of throttling mechanism. Both valves and vanes are well known throttling devices.

The diagrammatic illustration of the Brayton cycle system is similar to that shown in FIG. 1, but the cooling of the air is effected, as described above, in a refrigerant system involving an evaporator and condenser.

From the Brayton cycle air compressor the compressed air enters a recuperator together with condensate water from the refrigerant evaporator. In this recuperator, which, as shown, receives hot exhaust gases from the Brayton cycle expander, the temperature is above the boiling point of the condensate water and thus transforms it into water vapor, which, mixed with the heated compressed air, then passes through the combustion chamber and thence to the inlet of the Brayton cycle expander, in which it expands and the hot exhaust gases flow through the recuperator, as has been described. The Brayton cycle air compressor is shown as driven from the Brayton cycle expander from which a shaft extends, shown as broken, to any desired device for utilizing the power. In FIG. 1 this shaft is shown as going to a generator but in FIG. 3, in order to conserve drawing space, the generator is not shown.

FIG. 1 will next be described and will be followed by the computation for operation at maximum efficiency. These computations are more clearly understandable after a description of the actual machine elements.

On this figure air at ambient temperature T enters the air cooler at the point marked Air Intake, and is cooled to a temperature T which is at least 50F. below the ambient temperature at which it is intended to operate the engine. The cooled air enters the air compressor, in which it is compressed through the compression ratio r The ratio r is calculated for optimum operating conditions below in the section setting forth computations after description of the drawing. The compressed air enters the regenerator at a discharge temperature whichis determined by r and suction temperature T In the regenerator the air is heated by heat exchange with the exhaust from the Brayton cycle expander and passes into a conventional Brayton cycle combustion chamber. In this chamber fuel is burned and the temperature of the compressed air is raised further to T which is the maximum temperature that the-materials of the expander can withstand. The maximum permissible level of T is in no sense changed by the present invention.

The compressor is driven by the expander. The difference in the work which is produced by the expander and the work which is required by the compressor constitutes the net work output of the Brayton cycle. This is symbolized on the drawing by the power output shaft being connected to and driving an electric generator.

The expander exhaust gases go to a regenerator, which they leave at a temperature T T is determined by the discharge temperature of the compressed air and by the temperature differential AT indicated. The exhaust gases then pass through a refrigerant heater in which pump 2 keeps the refrigerant liquid at a sufficient pressure so that it does not boil. The amount of liquid which goes to the heater is determined by the adjustment of valves 5 and 6. In the heater the liquid refrigerant is heated up to temperature T minus the small temperature differential AT which is required for heat exchange. The exhaust gases then are exhausted as indicated, ordinarily at ambient temperature plus the same small temperature differential AT The hot refrigerant liquid flows from the refrigerant heater into a suitably insulated hot storage tank 1.

From time to time valve 4 is opened, and a batch of hot liquid is transferred from hot storage .tank 1 to hot flash tank 3. The liquid holding capacity of hot storage tank 1 is sufficiently greater than that of hot flash tank 3 to permit substantially continuous operation. The drawing is diagrammatic, so only a single hot flash tank is shown, but multiple tanks can be used, if desired.

In not flash tank 3 the heated refrigerant liquid, initially under such pressure as may be needed to prevent boiling in the refrigerant heater, flashes at decreasing temperatures and pressures until it reaches a minimum temperature and pressure, normally about ambient temperature. Valve 10 then is opened, and the remaining unvaporized liquid is permitted to flow into ambient storage tank 11.

Three refrigerant expanders 7, 8 and 9 constitute the power generating portion of the refrigeration system. The pattern of flow through the expanders is controlled by valves 12, 13, 14, 15, 16, 17, 18 and 19. At first, when the vapor in the hot flash tank is at maximum temperature and pressure, valves l2, l4, l7 and 19 are opened, and valves 13, 15, 16 and 18 are closed. As a result, refrigerant vapor passes in series through expanders 7, 8 and 9. These expanders drive corresponding refrigeration compressors 21, 22 and 23. This is symbolized on the drawing as a common shaft connecting expander 7 and compressor 23, a common shaft connecting expander 8 and compressor 22, and a common shaft connecting expander 9 and compressor 21.

At the start, the temperature and pressure in hot flash tank 3 is at a maximum and the flash vapor passes through expanders7, 8 and 9 in series. At the same time, the pressure and temperature in cold flash tank 30 is at a maximum, and the load on refrigeration compressors 21, 22 and 23 is at a minimum. The pattern of flow through these compressors is controlled by valves 20, 24, 25, 26, 27, 28, 29 and 33. At the start the three compressors operate in .parallel, valves 20, 24, 26, 27, 29 and 33 being open, and valves 25 and 28 being closed. The load on compressors 21, 22 and 23 increases as the temperature and pressure of the refrigerant in cold flash tank 30 drops. When expanders 7, 8 and 9 can no longer produce sufficient power to drive the compressors, valves 17, 19, 20 and 26 are closed. This has the effect of cutting off expander 9 and compressor 21; and now expanders 7 and 8 in series drive compressors 22 and 23 in parallel.

After a further lapse of time, the pressure and temperature of the refrigerant in hot flash tank 3 and of the refrigerant in cold flash tank 30 drops. When the load on compressors 22 and 23 increases and the power output of expanders 7 and 9 decreases to the point where the expanders can no longer drive the compressors,

valves 13 and 15 are opened and valve 14 is closed.

Now expanders 7 and 8 in parallel drive compressors 22 and 23 in parallel.

After a further drop in the temperature and pressure of the refrigerant in tanks 3 and 30, valves 27 and 29 are closed and valve 28 is opened. This results in two expanders, 7 and 8, in parallel driving two compressors, 22 and 23, in series.

When the temperatures and pressures in tanks 3 and 30 have dropped still further, valves 16, 19, 20 and 25 are opened and valve 24 is closed. Now the three expanders 7, 8 and 9 operate in parallel to drive the three compressors 21, 22 and 23 in series. It will be noted that during the whole operation exhaust vapors from the expanders and compressed vapors from the compressors flow into a conventional water cooled refrigeration condenser 34, where the vapors are condensed at practically ambient temperature. The condensate is discharged into ambient storage tank 11. When expanders 7, 8 and 9 no longer have sufficient power to drive compressors 21, 22 and 23, valve .10 is opened, and the unvaporized liquid in hot flash tank 3, now at substantially ambient temperature, also is discharged into ambient storage tank 11. The unvaporized cold liquid in cold flash tank 30 is discharged into cold storage tank 31 through valve 35. As in the case of hot storage tank 1, cold storage tank 31 should have sufficient capacity so that continuous operation is possible.

In the meantime pump 32 continuously has been pumping cold refrigerant liquid from cold storage tank 31 through the air cooler, which has been mentioned above. Flow of the cold liquid is controlled by valve 36. It will be seen from the drawing that the refrigerant liquid leaves the air cooler at substantially ambient temperature and flows into ambient storage tank 11, in which it is joined by the condensate which is formed in refrigeration condenser 34. Valves l and 35 now are closed and valves 4 and 6 are opened. A new batch of refrigerant liquid from hot storage tank 1 thus'is introduced into hot flash tank 3 and a new batch of ambient temperature liquid thus is introduced into cold flash tank 30. Therefrigeration cycle then is repeated. The system is self-regulating. If temperature T at the inlet of the air compressor tends to increase, the temperature of the compressed air entering the recuperator also increases and so, likewise, does T This, in turn, heats the refrigerant liquid to a higher temperature. The flashing of this hotter liquid in hot flash tank 3 produces more power which in turn reduces the temperature of the refrigerant in cold storage tank 31 and lowers T If T; tends to decrease, the process is reversed. This self-regulation is an advantage when the Treadwell System of refrigeration is combined with the recuperated Brayton cycle.

The preferred embodiment shown by the drawing utilizes all of the advantages of a full Treadwell System and represents a preferred modification, but the invention is not limited to using all of the advantages, and may use only part of them.

The following computations are made in conjunction with the combination of the cooled inlet Brayton cycle has been stated, as a refrigeration system forms the claimed subject matter of the parent application above referred to. The computations set forth quantities and values of certain of the quantities on the drawing, such as, for example, r The computations, however, in part are applicable to other refrigeration systems and are not all necessarily limited to the combination with the Treadwell system refrigeration.

The heat value of the refrigeration work (W which is required to cool a process stream (for example, a lb. .mol of gas, typically air) from ambient temperature (T,,) to some chosen lower temperature (T is 2 In A- M C (TS-ATE) where C the molal specific heat of the gas (for air about 7.0); T the chosen condensing temperature;

AT the chosen temperature difference between the with the Treadwell system of refrigeration, which, as l 5 318 BTU/lb. mol.

For any chosen refrigeration system a coefficient of performance C (at a particular T can be calculated, where C is the ratio of W (in heat units) to the heat removed from a lb. mol of gas when the gas is cooled from ambient temperature T, to a chosen lower temperature T Mathematically C [W,;]/[ Cp( T T and for the assumed set of conditions,

C R [3811/ [71)(5 60 40 0) 034 BTU of work required for each BTU which is removed in cooling the gas. Although for the chosen system C is dependent to some extent on the values of Cp, T AT T and E it is strongly dependent on T As an illustration, when the values of Cp, T AT T and E; are the same as in the previous example, but T is 450R rather than 400R, C 0.254 rather than 0.34. v

In the heat recovery system of the present invention, the heat value of the recovery work (W which is produced from the heat released by a process stream (for example a pound-mol of gas, typically air) as the stream cools from a super-ambient temperature (T to ambient temperature (T is (T -AT W =C E (ll'l C where E -is the chosen efficiency of the refrigerant expander, and AT is the chosen temperature difference between the gas and the refrigerant.

As an example, assume that E 0.8, AT 50, T

580R, and that it is necessary that the system produce 381 BTU of work per lb. mol,

W 7.0 X 0.8 [(T,; 580 (In 1)] 381 BTU/lb.mo1,

from which T 950R.

For any chosen heat recovery system a coefficient of where C is the ratio of W (in heat units) to the heat which becomes available when a lb. mol of gas cools from an initial temperature T to ambient temperature T,,. Mathematically, C [W ]/[C T -T and for the assumed set of conditions,

which is produced from each BTU of available heat.

For the chosen system, C is somewhat dependent on the values of Cp, E AT T, and T but it is strongly dependent on T As an illustration, when the values of Cp, E AT T,, and T are the same as in the previous example, but T is 900 rather than 950R, C 0.114 rather than 0.14.

The ambient temperature power producing refrigerant liquid is used countercurrently to cool the gas which is discharged from a gas compressor, and the ambient temperature power producing refrigerant liquid is heated thereby. When the gas suction temperature and the gas compression ratio are suitably matched, the

1 heat of compression of the gas heats the refrigerant power liquid to a temperature which is high enough so that the power liquid provides all of the work which is needed to refrigerate the gas which is about to enter the gas compressor, and no external work is needed to opperformance C (at a particular T can be calculated,

15 erate the refrigeration cycle. As an alternate the refrigeration cycle portion of the system may be powered in whole or in part by an independent motor or steam turbine driver. When the heat of gas compression provides the work of refrigeration, T is equal to the gas compressor discharge temperature, and T (T /E (r l) T wherein E is the efficiency of the gas compressor; r is the gas compression ratio; and n is the numerical value of adiabatic exponent (k-l )/k (for air, k 1.4, and n 0.286). When for a desired compression ratio r it is desired to determine the matching T a trial T is'selected and a corresponding T is calculated from the preceding formula. The refrigeration work W which is required for a chosen refrigeration system to cool the air to the trial T is calculated by the method previously explained. This W is compared to the calculated heat recovery'work W which is produced by a chosen heat recovery system (using the calculated T which corresponds to the trial T A series of values of T is tried until the refrigeration work which is required for the trial T is equal to the heat recovery work which is produced when the corresponding calculated T is used.

As an example, assume that a compression ratio of 15.0 is desired and that the self-driven Treadwell System is to be used. Several values of T are tried, which finally converge on 400R, and as a check, this value for T together with the desired value of 15.0 for r is substituted in the previously given equation u (TS/EC) '0" sln substituting,

T (400/185) (15.0 l) 400, from which T 950R. It was previously shown that with the Treadwell System, when T 950R, the work which is produced by the heat recovery system supplies the work which is required by the refrigeration system when T 400R.

With this suction temperature the gas compression requires only 7 l .5 percent of the single stage adiabatic work which is required when suction is taken at 560R. in the prior art a compression ratio of l5.0 cannot be achieved in a single stage compressor with ambient temperature (560R) suction because the discharge temperature of l,330R (870F) is much too high, and because far too much compression work is consumed, so a compression ratio of this magnitude usually requires two expensive, intercooled stages of compression. However, when the Treadwell System is used to cool the suction gas to 400R, the same compression ratio of .0 is readily conducted in a single stage compressor which produces a discharge temperature of only 950R (490F), and at the same time the net work is less than the work which is required by the more expensive two stage compressor. If other less efficient refrigeration and heat recovery systems are used in place of the Treadwell System, more refrigeration work is needed to cool the suction to 400R, and less heat recovery work is produced from the T of 950R. Therefore, the heat recovered cannot provide refrigeration to a temperature as low as 400R, and the net gas compression work is greater.

When the Treadwell System is used to cool the gas which is about to enter a gas compressor, the subsequent work of adiabatic compression closely approximates the work of isothermal compression when the isothermal compression process is conducted at ambient temperature. In fact, if AT and AT are made infinitely small, and T is made the same as T when the Treadwell System is used to cool the suction gas the work of adiabatic compression exactly equals that of ambient temperature isothermal compression.

lsotherrnal compression requires the least amount of work because in theory the process is reversible thermodynamically. With adiabatic compression, the gas which is discharged from the compressor is at a higher temperature than the gas which enters the compressor, and the heat energy which is required to produce this increase in temperature is' provided at the expense of additional work energy which has been delivered to the compressor. The compressed gas is discharged from the compressor at a relatively low temperature level and its heat normally is wasted by being rejected to cooling water in an inter or an after cooler. The direct rejection of this heat to cooling water is a completely irreversible process thermodynamically. By contrast, in the heat recovery portion of the Treadwell System heat is also rejected to cooling water, but only after it has produced work in the refrigerant expander. As a result, in the Treadwell System, in theory the heat rejection is completely reversible thermodynamically. Similarly, in the refrigeration portion of the Treadwell System, in theory the heat rejection is completely reversible. When in theory the Treadwell System is used with an adiabatic gas compressor, the gas initially is at ambient temperature and after compression and heat recovery is also at ambient temperature; the refrigeration process is reversible; the heat recovery process is reversible; and the adiabatic compression process is reversible. Since the final temperature of the compressed gas is the same as its initial temperature, and since in theory all of the processes involved are reversible, in theory adiabatic compression using the Treadwell System is equivalent to isothermal compression.

It will be noted that when the compressed gas supplies the heat which furnishes the work which is required by the refrigeration system, the system is selfregulating. If the gas compressor discharge temperature rises, more heat is available, more work is developed and more refrigeration work is available to lower the temperature of the gas which is about to enter the compressor. Whenthis temperature is lowered, the temperature of the gas which is discharged from the compressor is in turn lowered. If the gas compressor discharge temperature falls, less heat is available, less work is developed and less refrigeration work is available, so there is an increase in the temperature of the gas which is about to enter the compressor, and this in crease in turn raises the temperature of the gas which is discharged from the compressor. This automatic selfregulation is an important operating advantage of this aspect of the present invention.

The refrigeration system also can be used to cool substances other than gas. In such a case heat from another source may be used to raise the temperature of the refrigerant power liquid to a level high enough so that it will provide all the work which is needed by the refrigerant compressors. However, work is saved to the ex tent that waste heat is furnishing at least some of the work for the refrigeration system, even though it may not be all of the work.

When the Treadwell System is used to cool the gas entering a gas compressor, the heat of gas compression need not be the only source of heat for the power pro- 17 ducing refrigerant liquid. There may be other sources, which further can increase the amount of self-driven refrigeration that can be produced, and this can permit a still lower gas compressor inlet temperature, with a still further saving in compressor work.

The combination of the Treadwell System with a Brayton cycle engine constitutes the optimum form of the present invention wherein the air which is about to enter the compressor of a recuperated Brayton cycle engine is refrigerated and all the work of refrigeration is provided by the heat which is recovered fromthe exhaust air which is'leaving the recuperator or regenerator of the same engine. According to the presentinvention, it has been discovered that the maximum Brayton cycle work is produced when r has an optimum value defined by optimum where E (r "/r,, (the expansion ratio r,. is smaller than the compression ratio r because of parasitic pressure losses in the system); E and n are as previously defined; E is the air expander efficiency; T is the air expander inlet temperature; and T is a' chosen suction temperature, which is usually selected for practical reasons, such as the cost and the performance of available refrigeration equipment.

It has been discovered according to the preferred embodiment of the present invention that if the improvement in performance is to be of practical significance,

18 haust air which is leaving the recuperator and is, therefore, 120F. (950 830).,This is approximately the difference between the temperature of the compressed air which is leaving the regenerator and the temperature of the exhaust air which is leaving the expander and is entering the regenerator. For the assumed r the assumed E and the assumed Ep, the exhaust air leaves the expander at a temperature of about 1,160R, so the compressed air is heated recuperatively to about 1,040R (1 160 120) and the cycle operates at a thercause the air expander inlet temperature (T the system pressure losses which fix Ep; the component efficiencies (E E E E the heat exchanger temperature approaches (AT AT the condensing temperature (T and the ambient temperature (T are all constants for any selected system, the regenerator temperature approach (AT is a function only of T and r Since T and r are related by the previously given (Eq. 1 for optimum r for maximum Brayton cycle work output, r is'a function only of AT Therefore, when for economic or other reasons a specific regener- T should be at least about 1 below the ambient air 30 ator temperature approach is chosen, this choice also temperature that is ordinarily encountered. It has further been discovered that satisfactory results can be achieved over a range of from 10 percent greater to 10 I "EYE TH (Eq. 2), r

[ETEPTT 4ETE determines the optimum r which is required for maximum work output. 7

This optimum r is given by the equation.

percentsmaller than the r actually calculated.

For any chosen T there is a unique value of r at which the work produced in a Brayton cycle is a maximum. The net work output, i.e. the Brayton cycle work less the refrigeration work, depends, of course, on the efficiency of the refrigeration system which is chosen, but once the refrigeration system is chosen, at the chosen T, the net work output is a maximum at the same unique value of r at which (for the same T the Brayton cycle work output is a maximum.

It was shown previously that when the Treadwell System is used at the assumed conditions, the exhaust air must enter the heat recovery system at 950R in order for the heat recovery system to provide the refrigeration work .which is needed when the refrigeration system cools from 560R to 400R the air which is about to enter the air compressor. It wasalso shown previously that for maximum Brayton cycle work output the optimum r (E E E T /T 1.

As an example, assume that E 0.85; E 0.87; E, (1.05)" 1.014; T 1,960R (1,500F); and T 400R,

(0.85 X 0.87 X 1.014 X l960)/(400) 1.915 =r With this r E and T the air compressor discharge wherein E Ep, T T and E are as previously defined; AT is the chosen regenerator temperature approach; C is the coefficient of performance of the .chosen refrigeration system at the T; at which it operates; and C is the coefficient of performance of the chosen heat recovery system at the T at which it operates. lt is to be noted that although T does not appear explicitly in the equation, it is inherent in the calculaculated for this T and the coefficient of performance temperature (T is about 830R. The regenerator 6 entering the regeneratorand the temperature of the ex- C); of the chosen refrigeration system is calculated for the same trial T The values for T E AT C C T and E are substituted in Equation 2, and the resulting r is compared with the trial r Eq. 1. If this resulting 'r is not the same as the trial r a new trial r is calculated from a new trial T a new C and a new C are calculated, and a new resulting r is calculated. This procedure is repeated until the calculated resulting r is the same as the calculated trial r With a regenerator temperature approach of F (requiring an extremely large and very expensive regenerator), by calculation the suction temperature is 442R, the optimum r is 1.825, the cycle thermal efficiency is 42.8 percent, and the power production is 7.7 BTU of work for each cu. ft. of air displaced by the compressor, i.e. for each cu. ft. of compressor volumetric capacity. Other things being equal, the cost of a compressor is related to its volumetric capacity, and the cost of the compressor is a substantial part of the cost of a Brayton cycle engine. The work produced per unit of compressor capacity is therefore, a measure of the cost of the equipment used to produce power in a Brayton cycle engine.

With a regenerator temperature approach of 100F., the suction temperature is 407R, the optimum r is 1.898, the cycle thermal efficiency is 40.7 percent, and the power production is'8.6 BTU of work per cu. ft. of compressor capacity.

With a recuperator temperature approach of 120F., the suction temperature is 400F., th optimum r is 1.915, the cycle thermal efficiency is 40.2 percent, and the power production is 8.96 BTU of work per cu. ft. of compressor capacity. 7

With a regenerator temperature approach of 150F.,

the suction temperature is 390R, the optimum r is 1.94, the cycle thermal efficiency is 39.7 percent, and the power production is 9.42 BTU of work per cu. ft. of compressor capacity.

When no regenerator is used, and the heat of the air which is leaving the air expander is used only to power the refrigeration system, the suction temperature is 343R, the optimum r is 2.07, the cycle thermal efficiency i537 percent, and the power production is 12.3 BTU of work per cu. ft. of compressor capacity.

When taking suction at 560R, with an r of 1.62 and a regenerator temperature approachof 150, in the prior art a recuperated uncooled Brayton cycle engine produces 3.94 BTU of work per cu. ft. of compressor capacity, at a thermal efficiency of about 29.2 percent. The regenerator of this engine exhausts at about 1,120R, and when this exhaust heat is used to make 50 psig steam in a waste heat boiler, the steam produces in an expensive separate steam turbine about 1.38 BTU of additional work, for a total of 5.32 BTU for each cu.

ft. of capacity of the air compressor of the Brayton cycle engine. The combined cycle thermal efficiency is about 39 percent.

With the same regenerator temperature approach of 150, an engine designed in accordance with the present invention has a suction temperature of 390R, operates with a cycle thermal efficiency of about 39.7 percent, and produces a net work output of about 9.42 BTU per cu. ft. of compressor capacity, which is about 2.39'times that of the uncooled standard recuperated cycle engine and about 1.77 times that of the uncooled combined recuperated cycle engine. It is to be noted that this 9.42 BTU work output is produced by the Brayton cycle engine alone, and that there is no need for an expensive separate power producing steam turblue.-

All of the preceding Brayton cycle examples have been based on using the Treadwell System, but with a Brayton cycle it is possible to use other, less efficient combinations of self-driven refrigeration-heat-recovery systems. For example, a single or a multistage refrigeration system can be powered by a heat recovery system employing a single or a multistage boiler. For a chosen 20 heat recovery system and for a chosen refrigeration system it is necessary to give due consideration to all the pertinent factors, such as the heat exchanger temperature approaches, the component efficiencies, the number of stages, and the like, and to calculate for each specific T its coefficient of performance C and for each specific T its coefficient of performance C The previously described procedure using a trial T is then employed to determine the optimum r for a chosen nac- The preceding Eq. 2 for r is, therefore, applicable regardless of the type of refrigeration system or of the type of heat recovery system which is used.

The thermal efficiency of a prior art Brayton cycle engine suffers considerably if it becomes necessary to operate the engine at a reduced capacity. For example, the previously described prior art engine, with a regenerator temperature approach of 150, an r of 1.62, and a suction temperature of 560R (ambient temperature 560R), operates at a design point thermal efficiency of 29.2 percent. When called on to operate at 42 percent of its design point capacity, the same engine operates at a thermal efficiency of 17.8 percent, which is only 61 percent of its design point efficiency. This is because the most practical way to reduce the capacity of a prior art engine is to lower its firing temperature, which simultaneously lowers its Carnot cycle efficiency as well. In this example, at the design point the firing temperature is 1,500F., but at 42 percent capacity the firing temperature is only 1,090F.

An engine built in accordance with the present invention, with a regenerator temperature approach of 150, an r of 1.94, and a suction temperature of 390R (ambient temperature 560R), operates at a design point thermal efficiency of 39.7 percent. When this engine is called on to operate at 42 percent of its design point capacity, it operates at a thermal efficiency of 29.5 percent, which is almost percent of its design point efficiency. In this case, the capacity is reduced by allowing the level of suction refrigeration to rise to 560R. The engine capacity is easily varied between 42 percent and percent of its design capacity by adjusting the temperature level to which the suction air is cooled. (This adjustment can readily be made in various ways, for example, by throttling the flow of cooling water to the refrigeration condenser.) When the temperature of the suction air is raised, the engine capacity is reduced. The firing temperature remains at 1,500F. for this entire range'of engine capacities. If the firing temperature is lowered, the engine capacity can be reduced to even less than 42 percent of design capacity, at the penalty of a somewhat lowered thermal efficiency.

FIG. 2 illustrates the removal of contaminants from dirty fuel, producing clean fuel for the Brayton cycle engine. Some coals have contaminants of sulfur and particulates, and some oils have sulfur and a few also have small amounts of vanadium and nickel. The drawing is diagrammatic as the particular design of equipment used is not changed by the'present invention.

Dirty fuel,-either solid or liquid, is partially oxidized in a reactor 36. This reactor receives compressed air at about 100 psig, heated to 800F., through the pipe 37. The amount of air is restricted so that partial oxidation takes place. SomelOO psig steam is also introduced through the pipe 43 to aid in the gasification or partial oxidation of the fuel. A product gas at about 2,350F. enters the waste heat boiler 38 after having been mixed with a much larger stream, 10 mols to l, of cooler product gas from the pipe 47. This results in lowering the temperature of the mixture to about 850F., as indicated, a temperature which is sufficiently low so that hydrogen sulfide does not corrode the boiler surfaces.

Some steam is bled out of the turbine at the low pressure stage and is recycled through the pipe 43 to the reactor, as has been described. The process gases leave boiler 38 at about 700F., pass through the pipe 47, are blown by the blower 39 through a dust collector 41 and then are split, about 1 mol going to a boiler feedwater heater 45,- which supplies the boiler with feedwater through the pipe 46, and the larger portion about 10 -mols, passing on through the pipe 47 to quench the high temperature product gases entering the boiler, as has been described. While the volume recirculated is very large compared to the volume produced by partial oxidation, this recirculation or recycling constitutes a circulating load, and about 1 mol of product gases finally are cooled in the feedwater heater 45. There is no loss in heat as the sensible heat of the large volume of recirculated quenched gases gives up this sensible heat in the boiler 38 without significant loss.

The exhaust from the steam turbine 42 is condensed in the condenser 49 and pumped by the pump 50 through the feedwater heater 45 back to the boiler 38. Makeup water to compensate for the steam used in the partial oxidation is introduced into the suction of pump 50 in the conventional manner. This additional water, which though needed is not a feature of the present invention, is, therefore, not specifically shown on the drawing.

In the feedwater heater '45 the 750 product gas is cooled to about 150F. As in many places on the drawing, thisis an approximate temperature and is symbol- I ized by the :t symbol. The 15091 product gases are scrubbed in particulate scrubber 51, where solid mate-' of course the plant 52 maybe omitted. The purified gas passes through the pipe 53 to recuperator 54, which is heated by exhaust gases from the Brayton cycle'expander 50. The gas is heated up to about 800F. and the expander exhaust is cooled down to about 250F., passing out through the exhaust pipe 60.

In parallel with the recuperator 54 is a recuperator S9, which is also fed with a portion of the hot exhaust gases from the expander 57. Through this recuperator, air compressed in the air compressor 22 enters through pipe 48 and is heated up to about 800F. The expander 56, as is normal, produces power in excess of that required by the compressor 57 and this additional power is obtained as useful work, symbolized by the generator 58. The hot compressed air stream is split, part of it going to the partial oxidation reactor 36 through the pipe 37, as has been described, and part of it into the Brayton cycle combustor 55, where it burns with the fuel .gases preheated by the recuperator 54.

It will be seen that the major portion of the energy in the exhaust gases from the expander 56 is effectively used in preheating air and fuel for the Brayton cycle combustor and partial oxidation reactor. The only significant energy losses are in the exhaust pipe 60, which exhausts at a very much lower temperature than from an ordinary Brayton cycle expander, and a small amount lost in the particulate scrubber 51. Steam at a useful pressure and temperature is produced economically by the waste heat boiler 38 from the large volume of quenched process gases, and the boiler operates reliably since the inlet process gas temperature is brought down to a low enough figure to prevent damage to the heating surfaces of the boiler. This quenching, as has been described, does not result in any loss of heat because the sensible heat of the large volume of gases going through the boiler is practically all effectively utilized in raising steam.

FIG. 2 illustrates the combination of recycling a large amount of cooled partially oxidized gases in order'to bring down the temperature in the boiler 41. With certain contaminants the relatively coolsurfaces in the waste heat boiler 41 can tolerate a higher temperature, and where this is possible, part or all of the recycling of the cooled gases may be omitted.

FIG. 2 illustrates the combination of cleaning dirty fuel with an open Brayton cycle system. It is particularly useful with such a system when refrigerated suction of the present invention is used. However, the same partial oxidation and cleaning of the dirty fuel may be used with a closed Braytonv cycle.

I claim:

1. An open Brayton cycle comprising, in combination, air compression, combustion, and turbo expansion of the combustion gases, refrigerating the air at the inlet of compression, the refrigeration including refrigerant evaporation and refrigerant condensation in heat exchange relation with ambient air, and controlling refrigeration to maintainconstant a temperature differential between the refrigerant at evaporation and condensation.

2. A cycle according to claim 1 in which the refrigeration of the air being compressed is to a temperature at which water vapor condenses, vaporizing the water at a pressure at least as high as the pressure of the compressed air, and introducing the vaporized water into the Brayton cycle between air compression and turbo expansion.

3. A cycle according to claim 2 in which the water condensation is in the presence of an aqueous solution of a freezing point depressant.

4. A cycle according to claim 2 in which the water is condensed in the presence of a water vapor pressure depressant.

5. A cycle according to claim 1 in which the control of the refrigeration system is by variable throttling in the suction of the refrigerant compression.

6. A cycle according to claim 5 in which the refrigeration of the air being compressed is to a temperature at which water vapor condenses, vaporizing thewater at a pressure at least as high as the pressure of the compressed air, and introducing the vaporized water into which'apparatus includes in the Brayton cycle an air compressor, a combustion device into which compressed air passes, means for introducing fuel into said combustion device to burn with the compressed air, a turbo expander, and means for connecting the combus- 24 tion device to the inlet of the turbo expander, a refrigerating means for refrigerating air entering the air compressor, which refrigerating means includes a refrigerant evaporator in heat exchanging relation with the compressor air inlet, a refrigerant compressor and a refrigerant condenser in heat exchange relation with ambient air, and means for controlling the refrigeration to produce a substantially predetermined V constant evaporator-condensor temperature differential.

10. An apparatus according to claim 9 in which the control of refrigeration is by throttling means on the suction of the refrigerant compressor.

Claims (10)

1. An open Brayton cycle comprising, in combination, air compression, combustion, and turbo expansion of the combustion gases, refrigerating the air at the inlet of compression, the refrigeration including refrigerant evaporation and refrigerant condensation in heat exchange relation with ambient air, and controlling refrigeration to maintain constant a temperature differential between the refrigerant at evaporation and condensation.
2. A cycle according to claim 1 in which the refrigeration of the air being compressed is to a temperature at which water vapor condenses, vaporizing the water at a pressure at least as high as the pressure of the compressed air, and introducing the vaporized water into the Brayton cycle between air compression and turbo expansion.
3. A cycle according to claim 2 in which the water condensation is in the presence of an aqueous solution of a freezing point depressant.
4. A cycle according to claim 2 in which the water is condensed in the presence of a water vapor pressure depressant.
5. A cycle according to claim 1 in which the control of the refrigeration system is by variable throttling in the suction of the refrigerant compression.
6. A cycle according to claim 5 in which the refrigeration of the air being compressed is to a temperature at which water vapor condenses, vaporizing the water at a pressure at least as high as the pressure of the compressed air, and introducing the vaporized water into the Brayton cycle between air compression and turbo expansion.
7. A cycle according to claim 6 in which the water condensation is in the presence of an aqueous solution of a freezing point depressant.
8. A cycle according to claim 6 in which the water is condensed in the presence of a water vapor pressure depressant.
9. An apparatus for carrying out a Brayton cycle, which apparatus includes in the Brayton cycle an air compressor, a combustion device into which compressed air passes, means for introducing fuel into said combustion device to burn with the compressed air, a turbo expander, and meaNs for connecting the combustion device to the inlet of the turbo expander, a refrigerating means for refrigerating air entering the air compressor, which refrigerating means includes a refrigerant evaporator in heat exchanging relation with the compressor air inlet, a refrigerant compressor and a refrigerant condenser in heat exchange relation with ambient air, and means for controlling the refrigeration to produce a substantially predetermined constant evaporator-condensor temperature differential.
10. An apparatus according to claim 9 in which the control of refrigeration is by throttling means on the suction of the refrigerant compressor.
US3788066D 1970-05-05 1971-09-14 Refrigerated intake brayton cycle system Expired - Lifetime US3788066A (en)

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US5353585A (en) * 1992-03-03 1994-10-11 Michael Munk Controlled fog injection for internal combustion system
US5326254A (en) * 1993-02-26 1994-07-05 Michael Munk Fog conditioned flue gas recirculation for burner-containing apparatus
US5388395A (en) * 1993-04-27 1995-02-14 Air Products And Chemicals, Inc. Use of nitrogen from an air separation unit as gas turbine air compressor feed refrigerant to improve power output
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US5655373A (en) * 1994-09-28 1997-08-12 Kabushiki Kaisha Toshiba Gas turbine intake air cooling apparatus
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US5513488A (en) * 1994-12-19 1996-05-07 Foster Wheeler Development Corporation Power process utilizing humidified combusted air to gas turbine
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US6605190B1 (en) * 1997-02-14 2003-08-12 San Diego State University Foundation Staged optimal externally-controlled systems and method thereof
US6615585B2 (en) * 1998-03-24 2003-09-09 Mitsubishi Heavy Industries, Ltd. Intake-air cooling type gas turbine power equipment and combined power plant using same
US20020099476A1 (en) * 1998-04-02 2002-07-25 Hamrin Douglas A. Method and apparatus for indirect catalytic combustor preheating
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US9470149B2 (en) 2008-12-11 2016-10-18 General Electric Company Turbine inlet air heat pump-type system
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US9605890B2 (en) 2010-06-30 2017-03-28 Jmc Ventilation/Refrigeration, Llc Reverse cycle defrost method and apparatus
US9388817B1 (en) * 2011-03-24 2016-07-12 Sandia Corporation Preheating of fluid in a supercritical Brayton cycle power generation system at cold startup
US20150013364A1 (en) * 2012-01-25 2015-01-15 BSH Bosch und Siemens Hausgeräte GmbH Refrigeration device with a refrigeration compartment
US9410451B2 (en) * 2012-12-04 2016-08-09 General Electric Company Gas turbine engine with integrated bottoming cycle system
US20140150443A1 (en) * 2012-12-04 2014-06-05 General Electric Company Gas Turbine Engine with Integrated Bottoming Cycle System
US9380746B2 (en) 2013-03-15 2016-07-05 Storage Systems Northwest, Inc. Environmentally controlled storage facility for potatoes and other crops
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US10101092B2 (en) * 2014-08-22 2018-10-16 Peregrine Turbine Technologies, Llc Power generation system including multiple cores
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US20170233083A1 (en) * 2016-02-16 2017-08-17 The Boeing Company Thermal management systems and methods
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