US3668884A - Refrigeration system, heat recovery system, refrigerated gas compression system and brayton cycle system - Google Patents

Refrigeration system, heat recovery system, refrigerated gas compression system and brayton cycle system Download PDF

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US3668884A
US3668884A US34717A US3668884DA US3668884A US 3668884 A US3668884 A US 3668884A US 34717 A US34717 A US 34717A US 3668884D A US3668884D A US 3668884DA US 3668884 A US3668884 A US 3668884A
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refrigerant
compressors
liquid
hot
flash tank
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William H Nebgen
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02C—GAS-TURBINE PLANTS; AIR INTAKES FOR JET-PROPULSION PLANTS; CONTROLLING FUEL SUPPLY IN AIR-BREATHING JET-PROPULSION PLANTS
    • F02C7/00—Features, 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
    • F02C7/12—Cooling of plants
    • F02C7/14—Cooling of plants of fluids in the plant, e.g. lubricant or fuel
    • F02C7/141—Cooling of plants of fluids in the plant, e.g. lubricant or fuel of working fluid
    • F02C7/143—Cooling of plants of fluids in the plant, e.g. lubricant or fuel of working fluid before or between the compressor stages
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B11/00—Compression machines, plants or systems, using turbines, e.g. gas turbines
    • F25B11/02—Compression machines, plants or systems, using turbines, e.g. gas turbines as expanders
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02G—HOT GAS OR COMBUSTION-PRODUCT POSITIVE-DISPLACEMENT ENGINE PLANTS; USE OF WASTE HEAT OF COMBUSTION ENGINES; NOT OTHERWISE PROVIDED FOR
    • F02G2250/00—Special cycles or special engines
    • F02G2250/03—Brayton cycles
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—Component parts or details not otherwise provided for in this subclass
    • F25B2400/14—Power generation using energy from the expansion of the refrigerant
    • F25B2400/141—Power generation using energy from the expansion of the refrigerant the extracted power is not recycled back in the refrigerant circuit

Definitions

  • the above refrigerating system is preferably driven by a hot refrigerant liquid at progressively decreasing temperatures and pressures, thus cooling the remaining liquid to about ambient temperature.
  • the vapors flashed drive refrigerant turbines, each of which drives a refrigerant compressor, and after expansion the vapor is cooled and condensed at ambient temperature.
  • the pressure available to the refrigerant expanders is at a maximum.
  • the compression ratio required of the refrigerant compressors is at a minimum.
  • the pressure available to the expanders is at a minimum and the compression ratio required of the compressors is at a maximum.
  • the refrigerant expanders and compressors therefore are valved sequentially so that at the start of the cycle the expanders are in series and the compressors are in parallel and at the end of the cycle the expanders are in parallel and the compressors are in series.
  • air is compressed with refrigeration at compressor inlet at least 50 F. below ambient and to from 0.9 to 1.1 ofa calculated optimum compression ratio raised to the 0.286 power.
  • the efficiency of a refrigeration system is determined by the work which is required to remove the necessary quantity of heat from a process stream which it is desired to cool to some chosen temperature. When this heat is removed in stages of progressively lower temperatures, the work which is required is reduced in accordance with the number of stages employed; the greater the number of stages, the less is the work required and the greater is the efficiency of the refrigeration system.
  • the efficiency of a heat recovery system is determined by the work which it produces from the heat which it removes from a process stream when it cools the process stream through a chosen temperature range.
  • the work which is produced is increased in accordance with the number of stages employed; the greater the number of stages, the greater is the work produced, and the greater is the efficiency of the heat recovery system.
  • the gas which is about to enter a gas compressor is cooled by refrigeration, less work is required to compress the gas through a desired compression ratio, and if the inlet gas is cooled to'a suitable temperature the gas compression work which is saved is more than the refrigeration work which is needed to cool the gas.
  • the total work which is the sum of the cooled-suction gas compression work and the refrigeration work, is less than the uncooled-suction gas compression work alone.
  • the amount of work which is saved depends on the efficiency (the number of stages) of the refrigeration system and on the gas compression ratio; the more efficient the system and the greater the ratio, the greater the saving. With a given compression ratio, and when cooling to a given temperature, a two stage refrigeration system provides a greater saving than a single stage, and a three stage system provides a still greater saving.
  • a Brayton cycle engine air enters the engine at atmospheric pressure, is compressed, is heated 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 air expander and the work consumed by its air compressor.
  • the work produced by the air expander of a 5.4 ratio simple 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 100 F.) the work consumed by the air compressor is about 1.77 times the net work output.
  • the work output of this Brayton cycle engine increases because the compression ratio increases, and the expansion ratio increases accordingly; because the work produced by the air expander therefore increases; and also because the mass flow 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 100 F.
  • One aspect of the present invention concerns an improved refrigeration system.
  • cold liquid phase refrigerant is used countercurrently to cool a process stream from about ambient temperature to some chosen lower temperature.
  • the cold refrigerant liquid thereby is heated to about ambient temperature.
  • the liquid refrigerant is maintained under sufficient pressure so that it does not boil at the top temperature which it reaches as it countercurrently cools the process stream.
  • batch of cold liquid refrigerant is produced by permitting a batch of ambient temperature refrigerant liquid to flash in a flash tank. During its flashing, the refrigerant liquid boils at progressively lower pressures and temperatures. The vapor which evolves is removed continuously by refrigerant compressors and then is compressed, is cooled, and is condensed at about ambient temperature. The flashing results in cooling down that portion of the refrigerant liquid which does not flash, and when the liquid thus has been cooled down suffciently it is stored temporarily in a cold storage tank and subsequently it is used countercurrently to cool the process stream. Although the cold liquid is produced in batches, the cold storage tanks permits using it at a continuous rate to cool the process stream.
  • the heat recovery system ambient temperature refrigerant liquid is heated and is maintained under sufficient pressure so that it does not boil while it is being heated.
  • the hot liquid is stored temporarily in a hot storage tank.
  • a batch of the heated liquid is transferred from the hot storage tank to a hot flash tank, where it is permitted to flash at decreasing temperatures and pressures until finally it reaches a predetermined lower temperature, usually about ambient temperature.
  • the flash vapor passes through refrigerant expanders, each one driving a corresponding refrigerant compressor, as has been referred to above.
  • the refrigerant expander may drive its refrigerant compressor directly or it may drive the compressor through a constant or through a variable speed changing device.
  • the flash vapor which leaves the last expander is condensed at about ambient temperature.
  • the vapor coming off the hot flash tank is at its maximum temperature and pressure and can produce the maximum amount of expansion work.
  • the vapor coming off the cold flash tank is also at its maximum temperature and pressure and it requires the minimum amount of compression work. Therefore, at the beginning of the cycle the available overall expansion ratio is used in three expanders which are valved so that they operate in series, and the required compression ratio is provided by the paired three compressors which are valved so that they operate in parallel.
  • the vapor leaving the last refrigerant expander is cooled to substantially ambient temperature and is condensed to a liquid.
  • the refrigerant compressor discharge vapor likewise is cooled and condensed.
  • the ambient temperature condensed refrigerant liquid may be the feed both for the cold flash tank and for the waste heat recovery system. (it is, of course, possible, and in some cases it may be desirable, to use different fluids for the refrigeration cycle and for the work producing cycle. in this case two separate condensers are used.
  • the ambient temperature liquid which is left in the hot flash tank is transferred to the ambient storage tank; the cold liquid which is left in the cold flash tank is transferred to the cold storage tank.
  • a fresh batch of hot refrigerant liquid is transferred from the hot storage tank to the hot flash tank, a fresh batch of ambient temperature refrigerant liquid is transferred from the ambient storage tank to the cold flash tank, and the cycle is repeated.
  • the heat value of the refrigeration work 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 ER [T(' in where C P the molal specific heat of the gas (for air about 7.0); T the chosen condensing temperature; AT the chosen temperature difference between the gas which is leaving the cooler and the refrigerant liquid which is entering the cooler; and E the chosen efficiency of the refrigerant compressor.
  • a process stream for example, a lb. mol of gas, typically air
  • T ER
  • C P the molal specific heat of the gas (for air about 7.0)
  • T the chosen condensing temperature
  • AT chosen temperature difference between the gas which is leaving the cooler and the refrigerant liquid which is entering the cooler
  • E the chosen efficiency of the refrigerant compressor.
  • 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.
  • the gas suction temperature and the gas compression ratio are suitably matched, the 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 operate the refrigeration cycle.
  • T is equal to the gas compressor discharge temperature
  • E the efficiency of the gas compressor
  • r the gas compression ratio
  • n the numerical value of adiabatic exponent (k-l/k) (for air, k L4, and n 0.286).
  • 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 recover 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.
  • T (400/l 85) (15.0 I) 400 from which T 950 R. It was previously shown that with the Treadwell System, when 7",, 950 R., the work which is produced by the heat recovery system supplies the work which is required by the refrigeration system when T 400 R.
  • 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 A7 ⁇ , and A7 ⁇ , 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.
  • Isothermal compression requires the least amount of work because in theory the process is reversible thermodynamically.
  • 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.
  • 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.
  • the heat rejection is completely reversible thermodynamically.
  • the heat rejection is completely reversible.
  • 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.
  • the system when the compressed gas supplies the heat which furnishes the work which is required by the refrigeration system, the system is self-regulating. 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. When this 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 increase in turn raises the temperature of the gas which is discharged from the compressor. This automatic self-regulation 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.
  • 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.
  • work is saved to the extent 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.
  • the heat of gas compression need not be the only source of heat for the power producing 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 another preferred 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 from the exhaust air which is leaving the recuperator of the same engine.
  • the maximum Brayton cycle work is produced when r has an optimum value defined by optimum r [(E E E T IT H (Eq.
  • E is the air expander efficiency
  • T is the air expander inlet temperature
  • T is a chosen suction temperature, which is usually selected for practical reasons, such as the cost and the performance of available refrigeration equipment.
  • T should be at least about 50 F. below the ambient air 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 percent smaller than the r actually calculated.
  • 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.
  • E Ep, T T and E are as previously defined; AT is the chosen recuperator temperature approach; C is the coefficient of performance of the chosen refrigeration system at the T at which it operates; and C is the coefiicient of performance of the chosen heat recovery system at the T at which it operates. It is to be noted that although T does not appear explicitly in the equation, it is inherent in the calculation of C and C The preceding equation holds true only when the work of refrigeration is supplied by the heat which is available in the recuperator exhaust.
  • a trial T is selected, and from r (E E E T /T W relating r to T at trial optimum r is calculated.
  • the corresponding air compressor discharge temperature T then is calculated and to it is added the desired recuperator temperature approach (AT to give 7'
  • the coefiicient of performance C of the chosen heat recovery system is calculated for this T and the coefficient of performance C R 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.
  • 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.
  • the suction temperature is 407 R.
  • the optimum r is 1.898
  • the cycle thermal efficiency is 40.7 percent
  • the power production is 8.6 BTU of work per cu. ft. of compressor capacity.
  • the suction temperature is 400 F the optimum r5 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.
  • the suction temperature is 390 R.
  • the optimum r is 1.94
  • the cycle thermal efficiency is 39.7 percent
  • the power production is 9.42 BTU of work per cu. ft. of compressor capacity.
  • the suction temperature is 343 R.
  • the optimum r is 2.07
  • the cycle thermal efficiency is 37 percent
  • the power production is 12.3 BTU of work per cu. ft. of compressor capacity.
  • recuperated uncooled Brayton cycle engine When taking suction at 560 R., with an r of 1.62 and a recuperator temperature approach of 150, in the prior art a recuperated uncooled Brayton cycle engine produces 394 BTU of work per cu. ft. of compressor capacity, at a thermal efiiciency of about 29.2 percent.
  • the recuperator of this engine exhausts at about 1,120 R., 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.
  • an engine designed in accordance with the present invention has a suction temperature of 390 R., 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 turbine.
  • 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.
  • a single or a multi-stage refrigeration system can be powered by a heat recovery system employing a single or a multi-stage boiler.
  • 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 50 F. 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 calculated optimum compression ratio r
  • the compressed air enters the recuperator at a discharge temperature which is determined by this optimum r and suction temperature T
  • 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 the generator.
  • the expander exhaust gases go to a recuperator, 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 I
  • 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 (I) 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.
  • 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 l 1).
  • 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 (l2), (l3), (l4), (l5), (l6), (17), (18), and (19).
  • valves (l2), (14), (17) and (19) are opened, and valves (13 ⁇ ,(15), (l6) and(l8) are closed.
  • refrigerant vapor passes in series through (7), (8) and 9).
  • These expanders drive corresponding refrigeration compressors (21), (22) and (23).
  • 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.
  • valves (l7), (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.
  • 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.
  • 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).
  • 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 is discharged into cold storage tank (31) through valve
  • cold storage tank (31) should have sufficient capacity so that continuous operation is possible.
  • valve (36) Valves (l0) 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). The refrigeration cycle then is repeated.
  • the system is self-regulating.
  • the parallel-series flow pattern of the refrigeration compressors may be used with any type of driver, and is not restricted to waste heate powered expanders. However, where waste heat is available, it is desirable to use it to the maximum extent possible.
  • waste heat it is not essential that the same refrigerant which is used in the power producing cycle be used in the refrigeration cycle, but when the Treadwell System is combined with a Brayton cycle, it is ordinarily more convenient and economical to use the same refrigerant liquid both for power production and refrigeration.
  • a refrigerating system in which a volatile refrigerant is chilled by volatilizing a portion thereof to produce cold liquid refrigerant, the volatilization being at decreasing pressures, comprising in combination a flash tank for liquid refrigerant, a plurality of compressors, power sources for driving each compressor, each compressor having a suction inlet for a volatilized refrigerant and a compressed vapor outlet, valved conduits extending from the refrigerant flash tank at a level above liquid refrigerant therein to t e suction inlets of the compressors, and valved conduits connecting the compressed vapor outlet from at least one of the compressors to the suction inlet of another compressor, whereby upon actuation of the valves vapors from the refrigerant flash tank can be directed to the suctions of all of the compressors in parallel or to two or more compressors in series, a cooled compressed vapor condenser, a source of coolant to the condenser at a temperature sufficiently low to cond
  • each compressor is connected to its own power turbine to drive it, a source of hot liquid refrigerant, a hot liquid refrigerant flash tank, means for transferring a charge of hot refrigerant liquid to the flash tank, each power turbine having a turbine vapor inlet and outlet, valved conduits connecting the turbine inlets to the hot flash tank at a point above the hot refrigerant liquid level therein, valved conduits connecting power turbine outlets to power turbine inlets of another power turbine, means for actuating the valves in the conduits so that at first the power turbines are in series while the refrigerant compressors are in parallel, whereby the hot refrigerant liquid flashes to predetermined temperature and pressure, then connecting power turbines in parallel as compressors are con-nected in series, until finally all of the power turbines are in parallel and all of the compressors are in series until the hot refrigerant flash tank has flashed to a minimum temperature and pressure, valved conduits connecting power turbine exhausts to the compressed refrigerant
  • Refrigerating system in which the number of turbines and compressors is three each.

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US34717A 1970-05-05 1970-05-05 Refrigeration system, heat recovery system, refrigerated gas compression system and brayton cycle system Expired - Lifetime US3668884A (en)

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AU (1) AU2847071A (de)
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CA (1) CA939523A (de)
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Cited By (30)

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US3783614A (en) * 1972-02-10 1974-01-08 H Walker Turbine engine
US4019343A (en) * 1976-01-13 1977-04-26 Roberts Edward S Refrigeration system using enthalpy converting liquid turbines
US4144723A (en) * 1976-03-15 1979-03-20 General Atomic Company Power plant secondary coolant circuit
US4187694A (en) * 1978-11-21 1980-02-12 Midolo Lawrence L Binary working fluid air conditioning system
USRE30630E (en) * 1976-03-15 1981-06-02 General Atomic Company Power plant secondary coolant circuit
US4347714A (en) * 1980-07-25 1982-09-07 The Garrett Corporation Heat pump systems for residential use
US4347711A (en) * 1980-07-25 1982-09-07 The Garrett Corporation Heat-actuated space conditioning unit with bottoming cycle
US4424667A (en) 1982-06-07 1984-01-10 Fanning Arthur E Apparatus for increasing the efficiency of a gas turbine engine
US4445639A (en) * 1980-07-25 1984-05-01 The Garrett Corporation Heat pump systems for residential use
US5150585A (en) * 1991-04-17 1992-09-29 Stanley Markiewicz Energy recovery system for cold storage warehouse
US5586429A (en) * 1994-12-19 1996-12-24 Northern Research & Engineering Corporation Brayton cycle industrial air compressor
EP0846220A4 (de) * 1995-08-24 2000-03-22 Charles R Kohlenberger Verfahren und vorrichtung zum kühlen der einlassluft einer verbrennungsgasturbine
US6449934B1 (en) 1995-11-13 2002-09-17 Ransomes America Corporation Electric riding mower with motor generator set and noise abatement
US6519946B2 (en) * 2000-06-09 2003-02-18 The Japan Steel Works, Ltd. Cogeneration system using waste-heat gas generated in micro gas turbine
US20050217288A1 (en) * 2004-03-31 2005-10-06 Denso Corporation & Nippon Soken, Inc. System utilizing waste heat of internal combustion engine
US20070074535A1 (en) * 2005-09-30 2007-04-05 Sullair Corporation Cooling system for a rotary screw compressor
US20070095039A1 (en) * 2005-11-01 2007-05-03 Textron Inc. Modular power source for riding mower
US20080163642A1 (en) * 2005-03-15 2008-07-10 Masakazu Okamoto Refrigeration Apparatus
US20090192691A1 (en) * 2008-01-28 2009-07-30 Textron Inc. Turf Maintenance Vehicle All-Wheel Drive System
US20090284011A1 (en) * 2008-05-16 2009-11-19 Mcbride Thomas S Continuos-Absorption Turbine
US7854293B2 (en) 2007-02-20 2010-12-21 Textron Innovations Inc. Steering operated by linear electric device
US20120266607A1 (en) * 2011-04-25 2012-10-25 Denso Corporation Magneto-caloric effect type heat pump apparatus
CN102914107A (zh) * 2011-08-04 2013-02-06 特灵空调系统(中国)有限公司 制冷系统的热能回收控制方法
US20130312452A1 (en) * 2011-03-31 2013-11-28 Carrier Corporation Expander System
CN109026241A (zh) * 2018-08-30 2018-12-18 中国科学院工程热物理研究所 一种热泵压缩空气储能系统
US10168078B2 (en) 2013-05-02 2019-01-01 Mayekawa Mfg. Co., Ltd. Refrigeration system
US10208985B2 (en) * 2016-12-30 2019-02-19 Heatcraft Refrigeration Products Llc Flash tank pressure control for transcritical system with ejector(s)
CN116950739A (zh) * 2023-08-09 2023-10-27 中国电建集团重庆工程有限公司 布雷顿循环联合闪蒸循环发电系统及方法
US20240011685A1 (en) * 2022-07-05 2024-01-11 Heatcraft Refrigeration Products Llc Hot Gas Defrost Using a Work Recovery Device
US11879363B2 (en) * 2020-03-30 2024-01-23 Xuanhua Guo Combined system of intercooled recuperative gas turbine and refrigerant compound bottoming cycle

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US2511716A (en) * 1945-03-17 1950-06-13 Katzow Abram Heat operated compression refrigeration
US2991632A (en) * 1958-12-11 1961-07-11 John G Rogers Refrigeration system
US3074249A (en) * 1960-06-15 1963-01-22 Ray M Henderson Refrigeration system and apparatus having a heating cycle and a cooling cycle
US3306062A (en) * 1965-08-18 1967-02-28 William M Reid Refrigeration system

Cited By (41)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3783614A (en) * 1972-02-10 1974-01-08 H Walker Turbine engine
US4019343A (en) * 1976-01-13 1977-04-26 Roberts Edward S Refrigeration system using enthalpy converting liquid turbines
US4144723A (en) * 1976-03-15 1979-03-20 General Atomic Company Power plant secondary coolant circuit
USRE30630E (en) * 1976-03-15 1981-06-02 General Atomic Company Power plant secondary coolant circuit
US4187694A (en) * 1978-11-21 1980-02-12 Midolo Lawrence L Binary working fluid air conditioning system
US4445639A (en) * 1980-07-25 1984-05-01 The Garrett Corporation Heat pump systems for residential use
US4347711A (en) * 1980-07-25 1982-09-07 The Garrett Corporation Heat-actuated space conditioning unit with bottoming cycle
US4347714A (en) * 1980-07-25 1982-09-07 The Garrett Corporation Heat pump systems for residential use
US4424667A (en) 1982-06-07 1984-01-10 Fanning Arthur E Apparatus for increasing the efficiency of a gas turbine engine
US5150585A (en) * 1991-04-17 1992-09-29 Stanley Markiewicz Energy recovery system for cold storage warehouse
US5586429A (en) * 1994-12-19 1996-12-24 Northern Research & Engineering Corporation Brayton cycle industrial air compressor
EP0846220A4 (de) * 1995-08-24 2000-03-22 Charles R Kohlenberger Verfahren und vorrichtung zum kühlen der einlassluft einer verbrennungsgasturbine
US6449934B1 (en) 1995-11-13 2002-09-17 Ransomes America Corporation Electric riding mower with motor generator set and noise abatement
US6519946B2 (en) * 2000-06-09 2003-02-18 The Japan Steel Works, Ltd. Cogeneration system using waste-heat gas generated in micro gas turbine
US20050217288A1 (en) * 2004-03-31 2005-10-06 Denso Corporation & Nippon Soken, Inc. System utilizing waste heat of internal combustion engine
US7181919B2 (en) * 2004-03-31 2007-02-27 Denso Corporation System utilizing waste heat of internal combustion engine
US20080163642A1 (en) * 2005-03-15 2008-07-10 Masakazu Okamoto Refrigeration Apparatus
US7762099B2 (en) * 2005-03-15 2010-07-27 Daikin Industries, Ltd. Refrigeration apparatus
US20070074535A1 (en) * 2005-09-30 2007-04-05 Sullair Corporation Cooling system for a rotary screw compressor
US7334428B2 (en) * 2005-09-30 2008-02-26 Sullair Corporation Cooling system for a rotary screw compressor
US7392869B2 (en) 2005-11-01 2008-07-01 Textron Inc. Modular power source for riding mower
US20070095039A1 (en) * 2005-11-01 2007-05-03 Textron Inc. Modular power source for riding mower
US7854293B2 (en) 2007-02-20 2010-12-21 Textron Innovations Inc. Steering operated by linear electric device
US20090192691A1 (en) * 2008-01-28 2009-07-30 Textron Inc. Turf Maintenance Vehicle All-Wheel Drive System
US8521384B2 (en) 2008-01-28 2013-08-27 Textron Innovations Inc. Turf maintenance vehicle all-wheel drive system
US20090284011A1 (en) * 2008-05-16 2009-11-19 Mcbride Thomas S Continuos-Absorption Turbine
US9316419B2 (en) * 2011-03-31 2016-04-19 Carrier Corporation Expander system
US20130312452A1 (en) * 2011-03-31 2013-11-28 Carrier Corporation Expander System
US20120266607A1 (en) * 2011-04-25 2012-10-25 Denso Corporation Magneto-caloric effect type heat pump apparatus
US9534814B2 (en) * 2011-04-25 2017-01-03 Denso Corporation Magneto-caloric effect type heat pump apparatus
CN102914107B (zh) * 2011-08-04 2014-12-17 特灵空调系统(中国)有限公司 制冷系统的热能回收控制方法
CN102914107A (zh) * 2011-08-04 2013-02-06 特灵空调系统(中国)有限公司 制冷系统的热能回收控制方法
US10168078B2 (en) 2013-05-02 2019-01-01 Mayekawa Mfg. Co., Ltd. Refrigeration system
US10208985B2 (en) * 2016-12-30 2019-02-19 Heatcraft Refrigeration Products Llc Flash tank pressure control for transcritical system with ejector(s)
CN109026241A (zh) * 2018-08-30 2018-12-18 中国科学院工程热物理研究所 一种热泵压缩空气储能系统
CN109026241B (zh) * 2018-08-30 2023-12-08 中国科学院工程热物理研究所 一种热泵压缩空气储能系统
US11879363B2 (en) * 2020-03-30 2024-01-23 Xuanhua Guo Combined system of intercooled recuperative gas turbine and refrigerant compound bottoming cycle
US20240011685A1 (en) * 2022-07-05 2024-01-11 Heatcraft Refrigeration Products Llc Hot Gas Defrost Using a Work Recovery Device
US12163709B2 (en) * 2022-07-05 2024-12-10 Heatcraft Refrigeration Products Llc Hot gas defrost using a work recovery device
CN116950739A (zh) * 2023-08-09 2023-10-27 中国电建集团重庆工程有限公司 布雷顿循环联合闪蒸循环发电系统及方法
CN116950739B (zh) * 2023-08-09 2024-05-28 中国电建集团重庆工程有限公司 布雷顿循环联合闪蒸循环发电系统及方法

Also Published As

Publication number Publication date
FR2091249A5 (de) 1972-01-14
DE2122064A1 (de) 1971-12-09
ZA712173B (en) 1972-11-29
GB1349576A (en) 1974-04-03
ES390877A1 (es) 1973-07-01
BE766745A (fr) 1971-11-05
CA939523A (en) 1974-01-08
ATA386471A (de) 1975-06-15
NL7105785A (de) 1971-11-09
AU2847071A (en) 1972-11-09
IT942575B (it) 1973-04-02

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