US20070095097A1 - Thermal control system and method - Google Patents
Thermal control system and method Download PDFInfo
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- US20070095097A1 US20070095097A1 US11/591,465 US59146506A US2007095097A1 US 20070095097 A1 US20070095097 A1 US 20070095097A1 US 59146506 A US59146506 A US 59146506A US 2007095097 A1 US2007095097 A1 US 2007095097A1
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- 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
- F25B41/00—Fluid-circulation arrangements
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- 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—General 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
- F25B2400/01—Heaters
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- 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—General 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
- F25B2400/04—Refrigeration circuit bypassing means
- F25B2400/0403—Refrigeration circuit bypassing means for the condenser
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- 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—General 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
- F25B2400/04—Refrigeration circuit bypassing means
- F25B2400/0411—Refrigeration circuit bypassing means for the expansion valve or capillary tube
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- 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
- F25B40/00—Subcoolers, desuperheaters or superheaters
Definitions
- Cowans et al discloses a system and method of temperature control in which the temperature of a load is varied by thermal exchange with a saturated fluid directly, without requiring an intervening heat transfer fluid.
- the saturated fluid is a refrigerant used in pure gas phase, or pure liquid phase, or more often in a mixed or saturated phase.
- the vaporizable refrigerant is processed through substantially conventional compression and condensation steps as provided by commercially available equipment. However, compressed but not condensed refrigerant from the compressor is separately controlled, later to be mixed with condensed and selectively expanded refrigerant.
- the saturated refrigerant after mixing is at a pressure which determines its temperature, and its thermal energy is transferred directly with the thermal load which is to be controlled in temperature. Used in this way, the saturated fluid can provide a wide range of temperatures at the thermal load. However this range can be extended, because at one extreme the thermal load can be heated by using the hot gas phase alone, and at the other extreme the load can be chilled using only refrigerant on the condensed phase, after expansion. In this direct transfer system, temperature changes can be rapid and set points can be controlled very precisely. Equipment costs are substantially reduced because the system does not require use of an intermediate thermal transfer fluid, pumps, or a heat exchanger for thermal transfer fluid.
- Processor components in a refrigeration system must operate satisfactorily through all phases in the compression and heat exchange cycle, so that the refrigerant must be at proper temperatures and pressures so that different functions can be performed.
- the input to a compressor system should be free of liquid, and in a particular pressure range, or efficiency will be lost or the compressor damaged, or both. Maintaining efficiency throughout can present other problems that require solutions consistent with overall system requirements. For example, if the fluid is in a saturated state, the compressed gas component contributes relatively very little to chilling the heat load during thermal transfer. Thus when the load temperature is to be dropped to a minimum, the presence of the gas limits heat exchange efficiency.
- response capabilities of systems in accordance with the invention can be employed to meet the operative requirements of many different temperature control combinations. Where time is of the essence in bringing a thermal load to a target temperature, anticipatory manipulation of the controller can be useful.
- TDSF transfer direct saturated fluid
- a counter-current heat exchanger may be positioned in the flow path to the load to interchange thermal energy between the incoming input, and the out-going refrigerant passing from the load back to the suction input of the compressor. This energy interchange both increases the temperature level of the input to the load and reduces the temperature of the return fluid.
- the return fluid may thereafter be brought to a level compatible with the demands of compressor operation.
- For further heating the input to the load can be passed through a heater operated by the controller and included in the input stream to the load to thereby raise the temperature well above the compressor capability if desired.
- a further feature of the system is the incorporation of a computer controlled solenoid bypass between the hot gas shunt that leads to the mixing circuit, and the return line from the load to the suction input of the compressor.
- This bypass is operated to remove some of the flow from the hot gas line when full hot gas flow might tend to make the loop gain of the servo system unstable and/or reduce the amount of cooling available at temperatures within the desired range. The result introduces a time delay in cooling the load to enable closer control of temperature.
- Control circuit adaptations may be introduced to realize further benefits from the concept. Where fast reaction to commands requiring fast temperature changes are needed, temporary and short term commands can be utilized to shorten response times. If, for example, the active part of a thermal load comprises the surface of a wafer holding chuck in a semiconductor processing system, and the heat exchange region is physically spaced apart from the upper surface of the chuck, some time may be needed to bring the chuck surface to a specified temperature. By introducing control algorithms which bypass the need for accumulation of tracking data, the wafer can be much more rapidly stabilized at a target temperature, increasing yield rates and lowering costs.
- FIG. 1 is a block diagram representation of a temperature control system using direct transfer of saturated fluids in accordance with the invention
- FIG. 2 is a generalized schematic diagram of one form of a vapor separator such as may be used in the system of FIG. 1 ;
- FIG. 3 is a graphical representation of the improved control stability achieved by utilizing a supplemental hot gas bypass coupling as shown in FIG. 1 ;
- FIG. 4 is a graphical representation of high temperature response in a system utilizing a hot gas heater and heat exchanger as shown in FIG. 1 .
- FIG. 1 a temperature control system is shown utilizing a novel refrigeration/heating cycle in accordance with the above-referenced previously filed Cowans et al application Ser. No. 11/057,383.
- the present system integrally incorporates features in this context for achieving improved stability and/or efficiency when directly transferring thermal energy using saturated fluids.
- a compressor 10 which may use a conventional refrigerant such as R507, or a different refrigerant depending upon the application, feeds compressed hot gas into a condenser 12 which also is arranged as part of a pressure regulating system.
- the compressor 10 output is fed into a hot gas line 14 which is separately split to a condensed refrigerant line 16 , the flow proportions being controlled by a digital controller system 20 , as described in the previously filed application.
- the hot gas line 14 includes a proportional valve 22 operated by the controller 20 so as to provide a pressure modulated flow, responsive to system requirements.
- TXV thermal expansion valve
- the output flow from the mixing circuit 26 then cools or heats the thermal load 28 , such as a semiconductor tool, after which the refrigerant is transferred on a return line 30 back to the suction input to the compressor 10 .
- the return line 30 includes a serial accumulator chamber 32 , in which a heater 34 operated by the controller 20 restores the temperature level, as and if necessary.
- Operation of the system has, among other things, the advantage of providing a very wide potential temperature control range from hot to cold (e.g. from ⁇ 120° C. to ⁇ 60° C.), extremely fast temperature adjustments, and also high precision (e.g. ⁇ 1° C.).
- hot to cold e.g. from ⁇ 120° C. to ⁇ 60° C.
- high precision e.g. ⁇ 1° C.
- a shutoff valve 40 is disposed in series with the proportional valve 22 .
- the shutoff valve 40 is bypassed by a shunt valve 46 which is in parallel with it.
- the condensed refrigerant line 16 includes a shutoff valve 42 and a shunt valve 44 , which bypasses the TXV 24 , all valves being operated by the controller 20 .
- the condenser system includes a condenser heat exchanger 50 which is cooled by water from a conventional source 52 , although other cooling fluids may be used.
- the water flow rate is governed by a flow control system 54 , so as to maintain the output pressure from the compressor 10 .
- a bypass valve 56 is disposed in parallel with the coolant flow control.
- a subcooler heat exchanger 60 is disposed in the return line 30 leading to the suction input to the compressor 10 .
- the subcooler heat exchanger 60 operates as a counterflow device, cooling the outgoing flow from the condenser 12 with returning fluid, which in most modes will be expanded and cooled gases, directed back to the compressor 10 .
- shunt loop 62 about the subcooler heat exchanger 60 includes a desuperheater valve 64 responsive to a temperature sensor 66 at the suction input to the compressor 10 . If the input pressure to the compressor 10 falls too low the flow is augmented by opening the desuperheater valve 64 . In the shunt loop 62 , these flows are derived from a T-junction 68 at the condenser 12 output.
- the return line from the load 28 is passed through an accumulator 37 which includes a heater 34 operated by the controller 20 .
- the system can thus act in response to temperature signals provided from a temperature sensor 78 associated with the load 28 to restore or equalize the temperature of the fluid in the return line.
- a conduit to the TXV 24 from sensor bulb 74 in communication with the return line 30 is used for external equalization of the TXV 24 .
- crankcase pressure regulator valve 76 also known as a “close on rise” (COR) valve 76 .
- Flows at different points in the circulating loop must often be brought into predetermined pressure and temperature ranges for components to work properly.
- the compressor 10 input must be maintained above a selected pressure range. This is accomplished by a hot gas bypass valve 82 responsive to a pressure sensor 80 at the input to the compressor 10 .
- the hot gas bypass valve 82 feeds back a portion of the compressor 10 output flow to the suction input in the event the input pressure is too low.
- the system as thus far described operates as described and in practice validates the concept and its advantages, but it also possesses certain advantages and potentials not immediately evident.
- the system in the high temperature mode, the system can be operated with the proportional valve 22 alone providing temperature modulation, and with the refrigeration line 16 being shut off by the valve 42 .
- the input to the load 12 is then solely the high temperature gaseous flow, but the temperature of the input to the load 28 can be further raised to an even higher level, suitably compensating for anticipated major heat losses at the load 28 at these temperatures.
- a counterflow heat exchanger 86 and a serially coupled electrical heater 88 are disposed between the mixing circuit 26 and the load 28 .
- the input temperature to the load 28 is detected by a sensor 90 , so that actuating signals can be applied from the controller 20 to the heater 88 subsequent to the mixing circuit 26 .
- a heater 92 is provided in the hot gas line 14 prior to the mixing circuit 26 , to be energized by the controller 20 to provide further heating. Reverse flow back toward the control valves is blocked by a suitably placed check valve 94 .
- the counterflow heat exchanger 86 keeps the ? temperature level down to a predetermined range in the suction line to the compressor 10 .
- the sequence of temperature changes in this mode is shown graphically in FIG. 4 , wherein the hot gas temperature is successively increased from the level (A) provided by the full open proportional valve to the higher level (B) at the HEX 86 output and then the final highest level, from the heater 88 .
- the condensed refrigerant line 16 includes, subsequent to the TXV 24 , a check valve 98 and a vapor separator 100 , an example of which is seen in more detail in FIG. 2 .
- Vapor which is collected in the separator 100 is directed along a vapor line to the return line 30 to the compressor 10 .
- the liquid proportion is accentuated by the vapor separator 100 , and increases the efficiency of the compressor 10 .
- FIG. 2 shows a schematic cross-sectional diagram of the liquid separator 100 .
- Somewhat similar devices are commonly used in vapor-cycle refrigeration systems to separate high pressure refrigerant gas emerging from the compressor from oil mist carried along with the refrigerant but this use in a thermodynamic function is novel.
- Gas and liquid enter the separator body 150 at entry port 153 as indicated by arrow 154 on FIG. 2 .
- the separator 100 functions by using a finely divided metal wool barrier 156 placed in the path of the gas such that all the refrigerant must pass through the wool 156 toward an outlet. Droplets coalesce on the surfaces of the metal wool barrier 156 and descend, under the influence of gravity, to the bottom of the separator cavity.
- the basic TDSF system enables providing useful heat to maintain the load in the range of 90° C. to 120° C. by delivering high pressure gas to the load at temperatures that are sometimes well in excess of the required load temperature level.
- a flow of 200 grams/second of R507 gas requires that the gas be heated to about 28.5° C. or more above 120° C. In giving up heat to the load by cooling 28.5° C. this gas flow will bring the load to the target temperature. Somewhat more than this amount is needed to provide drive for the needed transfer of heat across whatever heat exchanger is used.
- a temperature of 120° C. is as high as a typical commercial compressor readily withstands, whereas the improvement of FIG. 1 can provide gas at temperatures approaching 200° C. if the structural members used can support such levels.
- this is achieved by using the counter-current HEX 86 together with the extra electrical heater 88 in the input path to the load 28 , after the mixer 26 .
- the system When the system is supplying temperatures less than about 60° C. the system functions substantially the same as does the prior system.
- the hot gas from the compressor 10 first provides its maximum level [(A) in FIG. 4 .] which is raised to a higher level by flowing through the countercurrent HEX [(B) in FIG. 4 ].
- the heater 88 is activated by electronic controller 20 to provide adequate heat to raise the temperature of the refrigerant to the desired final value [(C) in FIG. 5 ].
- the counter-current HEX 86 isolates the bulk of the TDSF system from any adverse effects of the high temperature because the fluid emerging on the return line from the counter-current HEX 86 will be not much hotter than the fluid emerging from the high pressure outlet of the compressor.
- the system includes a further improvement providing adequate control during times when the TDSF system is closely controlling the temperature of an object that is being temperature controlled by the TDSF.
- a full flow of gas from proportional valve 22 overwhelms the controller function if the entire flow is mixed with the flow from TXV 24 .
- This condition is illustrated by the graph of FIG. 3 , which shows the instability that exists when there is full hot gas flow.
- the effect of a small change in flow from valve 22 is then such as to change the total thermal output of the mixture to an excessive degree, and the system can tend to become unstable.
- the loop gain of the servo system which includes the combined output of the valves 22 and 24 is too high if the full flow from 22 is used to mix with the flow from valve 24 .
- the solution used is to employ a bypass line 103 including a hot gas bypass (solenoid) valve 104 as shown in FIG. 1 .
- the hot gas bypass valve 104 is responsive to the controller 20 . When close control is needed, the valve 104 is opened, allowing some of the gas output of proportional valve 22 to bypass directly to the input of compressor 10 , which has the effect of reducing flow through the cooled load 28 .
- a check valve 106 in the bypass line prevents any flow from the TXV 24 from being bypassed when the proportional valve 22 is closed.
- Control is enhanced because the overall loop gain of the control servo circuit is reduced and thus easier to control.
- TDSF systems can respond to needed temperature changes even faster than the electronic controllers, when the controllers have to store a series of readings before establishing reaching a steady state condition.
- the entry of a new set point can initiate a time consuming sequence in which, while transitioning to a new target value, a succession of readings are required.
- a TDSF system has a faster response it has been found useful to enter an artificial and temporary temperature reading into the controller. A new sequence of readings is not needed because previously taken temperature measurements are retained and the controller operates without interrupting the prior sequence. This enables final temperature adjustment of the saturated fluid much more rapidly.
- the artificial temperature input is used to compensate for thermal delays that are inherent in the design of a tool.
- the input temperature By altering the input temperature artificially in step-wise fashion before starting application of power, control of the chuck temperature is both more rapid and precise.
- Other empirically derived artificial inputs may be used in other situations, for start-up or shut-down sequences.
Abstract
Description
- This invention relies for priority on previously filed application Ser. No. 11/057,383 of Kenneth W. Cowans et al, filed Feb. 15, 2005 and entitled “Thermal Control System and Method”, and on
provisional application 60/733,078 filed Nov. 4, 2005 by Kenneth W. Cowans et al and entitled “Thermal Control System and Method”. - The above mentioned utility application of Cowans et al discloses a system and method of temperature control in which the temperature of a load is varied by thermal exchange with a saturated fluid directly, without requiring an intervening heat transfer fluid. Typically the saturated fluid is a refrigerant used in pure gas phase, or pure liquid phase, or more often in a mixed or saturated phase. In accordance with the invention, the vaporizable refrigerant is processed through substantially conventional compression and condensation steps as provided by commercially available equipment. However, compressed but not condensed refrigerant from the compressor is separately controlled, later to be mixed with condensed and selectively expanded refrigerant. The saturated refrigerant after mixing is at a pressure which determines its temperature, and its thermal energy is transferred directly with the thermal load which is to be controlled in temperature. Used in this way, the saturated fluid can provide a wide range of temperatures at the thermal load. However this range can be extended, because at one extreme the thermal load can be heated by using the hot gas phase alone, and at the other extreme the load can be chilled using only refrigerant on the condensed phase, after expansion. In this direct transfer system, temperature changes can be rapid and set points can be controlled very precisely. Equipment costs are substantially reduced because the system does not require use of an intermediate thermal transfer fluid, pumps, or a heat exchanger for thermal transfer fluid.
- The system and method present unique challenges, as well as possibilities. Processor components in a refrigeration system must operate satisfactorily through all phases in the compression and heat exchange cycle, so that the refrigerant must be at proper temperatures and pressures so that different functions can be performed. For example, the input to a compressor system should be free of liquid, and in a particular pressure range, or efficiency will be lost or the compressor damaged, or both. Maintaining efficiency throughout can present other problems that require solutions consistent with overall system requirements. For example, if the fluid is in a saturated state, the compressed gas component contributes relatively very little to chilling the heat load during thermal transfer. Thus when the load temperature is to be dropped to a minimum, the presence of the gas limits heat exchange efficiency. Another factor affecting efficiency occurs in a different temperature range, resulting from limitations on the energy of compression that can be applied to the refrigerant. When the refrigerant is to be used for heating, the compressor brings the hot gas to a given level, such as 120° C. However, if substantial heating energy is needed at the load, a much higher temperature level must be reached. The return flow, after heat exchange with the load, cannot however be at levels that disrupt the pressure/temperature/enthalpy balance needed with a vaporizable refrigerant. It is highly desirable to eliminate such problems without introducing thermodynamic conditions which affect the integrity of the refrigeration cycle.
- The response capabilities of systems in accordance with the invention can be employed to meet the operative requirements of many different temperature control combinations. Where time is of the essence in bringing a thermal load to a target temperature, anticipatory manipulation of the controller can be useful.
- In a transfer direct saturated fluid (TDSF) thermal control system, a number of local variations in control loops and components improve efficiency and stability across a range of load temperatures. Cooling efficiency, at very low temperatures, for example, is markedly improved at the load by extracting at least part of the vapor components in the condensed fluid after expansion before mixing with compressed gas. The lowered mass flow through the line then returning from the load to the compressor reduces the pressure drop in said return line. To this end, a vapor separator is disposed in the line transporting liquid/vapor product before the mixing junction, and the extracted vapor is fed back into the refrigerant returning from the load to compressor while a higher proportion of liquid is fed to the load. Thus heat transfer is more efficient without affecting the temperature of the mix.
- Other aspects of the invention are concerned with maintenance of efficiency and improvement of temperature limits when using the high temperature capability of the saturated fluid system. A counter-current heat exchanger may be positioned in the flow path to the load to interchange thermal energy between the incoming input, and the out-going refrigerant passing from the load back to the suction input of the compressor. This energy interchange both increases the temperature level of the input to the load and reduces the temperature of the return fluid. The return fluid may thereafter be brought to a level compatible with the demands of compressor operation. For further heating the input to the load can be passed through a heater operated by the controller and included in the input stream to the load to thereby raise the temperature well above the compressor capability if desired.
- A further feature of the system is the incorporation of a computer controlled solenoid bypass between the hot gas shunt that leads to the mixing circuit, and the return line from the load to the suction input of the compressor. This bypass is operated to remove some of the flow from the hot gas line when full hot gas flow might tend to make the loop gain of the servo system unstable and/or reduce the amount of cooling available at temperatures within the desired range. The result introduces a time delay in cooling the load to enable closer control of temperature.
- Control circuit adaptations may be introduced to realize further benefits from the concept. Where fast reaction to commands requiring fast temperature changes are needed, temporary and short term commands can be utilized to shorten response times. If, for example, the active part of a thermal load comprises the surface of a wafer holding chuck in a semiconductor processing system, and the heat exchange region is physically spaced apart from the upper surface of the chuck, some time may be needed to bring the chuck surface to a specified temperature. By introducing control algorithms which bypass the need for accumulation of tracking data, the wafer can be much more rapidly stabilized at a target temperature, increasing yield rates and lowering costs.
- A better understanding of the invention may be had by reference to the following description taken in conjunction with the accompanying drawings, in which:
-
FIG. 1 is a block diagram representation of a temperature control system using direct transfer of saturated fluids in accordance with the invention; -
FIG. 2 is a generalized schematic diagram of one form of a vapor separator such as may be used in the system ofFIG. 1 ; -
FIG. 3 is a graphical representation of the improved control stability achieved by utilizing a supplemental hot gas bypass coupling as shown inFIG. 1 ; and -
FIG. 4 is a graphical representation of high temperature response in a system utilizing a hot gas heater and heat exchanger as shown inFIG. 1 . - Referring now specifically to
FIG. 1 , a temperature control system is shown utilizing a novel refrigeration/heating cycle in accordance with the above-referenced previously filed Cowans et al application Ser. No. 11/057,383. The present system integrally incorporates features in this context for achieving improved stability and/or efficiency when directly transferring thermal energy using saturated fluids. - In
FIG. 1 , acompressor 10, which may use a conventional refrigerant such as R507, or a different refrigerant depending upon the application, feeds compressed hot gas into acondenser 12 which also is arranged as part of a pressure regulating system. Thecompressor 10 output is fed into ahot gas line 14 which is separately split to a condensedrefrigerant line 16, the flow proportions being controlled by adigital controller system 20, as described in the previously filed application. Thehot gas line 14 includes aproportional valve 22 operated by thecontroller 20 so as to provide a pressure modulated flow, responsive to system requirements. In theseparate line 16, condensed refrigerant flow is fed through and regulated by a thermal expansion valve (TXV) 24 which reduces the pressure of the condensed input, and expands the volume so as to lower the temperature, again in accordance with vapor-cycle operation and operating objectives. These two separate lines come together in amixing circuit 26 and as described in the prior case, in the principal range of operation the relative flows are adjusted so as to provide an output temperature determined by the pressure. - The output flow from the
mixing circuit 26 then cools or heats thethermal load 28, such as a semiconductor tool, after which the refrigerant is transferred on areturn line 30 back to the suction input to thecompressor 10. Thereturn line 30 includes aserial accumulator chamber 32, in which aheater 34 operated by thecontroller 20 restores the temperature level, as and if necessary. - Operation of the system has, among other things, the advantage of providing a very wide potential temperature control range from hot to cold (e.g. from ±120° C. to −60° C.), extremely fast temperature adjustments, and also high precision (e.g. ±1° C.). In addition, since no intermediate heat transfer system or medium is needed, these unique capabilities can be provided with substantial cost savings.
- However, the use of a saturated fluid in different phases (liquid, vapor, and saturated liquid/vapor phase), introduces a number of operative problems or conditions that should be accounted for to realize system potential more fully. In some applications it is desirable to effect rapid changes between operating modes at different temperature levels. For abrupt cessation of flow in the
hot gas line 14, ashutoff valve 40 is disposed in series with theproportional valve 22. For virtually immediate or assured full flow of hot gas, theshutoff valve 40 is bypassed by ashunt valve 46 which is in parallel with it. For rapid control of the condensed refrigerant flow, the condensedrefrigerant line 16 includes ashutoff valve 42 and ashunt valve 44, which bypasses theTXV 24, all valves being operated by thecontroller 20. The condenser system includes acondenser heat exchanger 50 which is cooled by water from aconventional source 52, although other cooling fluids may be used. In a configuration known in the prior art, the water flow rate is governed by aflow control system 54, so as to maintain the output pressure from thecompressor 10. To facilitate maximum cooling, abypass valve 56 is disposed in parallel with the coolant flow control. - Expedients are also used to improve system response and reliability in terms of thermodynamic efficiency. A
subcooler heat exchanger 60 is disposed in thereturn line 30 leading to the suction input to thecompressor 10. Thesubcooler heat exchanger 60 operates as a counterflow device, cooling the outgoing flow from thecondenser 12 with returning fluid, which in most modes will be expanded and cooled gases, directed back to thecompressor 10. In accordance with the W. W. Cowans Pat. No. 6,446,446 referenced in the predecessor parent application,shunt loop 62 about thesubcooler heat exchanger 60 includes adesuperheater valve 64 responsive to atemperature sensor 66 at the suction input to thecompressor 10. If the input pressure to thecompressor 10 falls too low the flow is augmented by opening thedesuperheater valve 64. In theshunt loop 62, these flows are derived from a T-junction 68 at thecondenser 12 output. - In order to preserve pressure and temperature balance in the closed loop compression/heat exchange system, the return line from the
load 28 is passed through an accumulator 37 which includes aheater 34 operated by thecontroller 20. The system can thus act in response to temperature signals provided from a temperature sensor 78 associated with theload 28 to restore or equalize the temperature of the fluid in the return line. Also, a conduit to theTXV 24 fromsensor bulb 74 in communication with thereturn line 30 is used for external equalization of theTXV 24. - Another feature cooperates with these elements and relationships to overcome different potential problems. If the pressure in the return suction line to the
compressor 10 becomes too high, it is automatically lowered by an included crankcase pressure regulator valve, also known as a “close on rise” (COR)valve 76. - Flows at different points in the circulating loop must often be brought into predetermined pressure and temperature ranges for components to work properly. For example, the
compressor 10 input must be maintained above a selected pressure range. This is accomplished by a hotgas bypass valve 82 responsive to a pressure sensor 80 at the input to thecompressor 10. The hotgas bypass valve 82 feeds back a portion of thecompressor 10 output flow to the suction input in the event the input pressure is too low. - The system as thus far described operates as described and in practice validates the concept and its advantages, but it also possesses certain advantages and potentials not immediately evident. For example, in the high temperature mode, the system can be operated with the
proportional valve 22 alone providing temperature modulation, and with therefrigeration line 16 being shut off by thevalve 42. The input to theload 12 is then solely the high temperature gaseous flow, but the temperature of the input to theload 28 can be further raised to an even higher level, suitably compensating for anticipated major heat losses at theload 28 at these temperatures. For this purpose, acounterflow heat exchanger 86 and a serially coupledelectrical heater 88 are disposed between the mixingcircuit 26 and theload 28. The input temperature to theload 28 is detected by asensor 90, so that actuating signals can be applied from thecontroller 20 to theheater 88 subsequent to the mixingcircuit 26. Separately, aheater 92 is provided in thehot gas line 14 prior to the mixingcircuit 26, to be energized by thecontroller 20 to provide further heating. Reverse flow back toward the control valves is blocked by a suitably placedcheck valve 94. Thecounterflow heat exchanger 86 keeps the ? temperature level down to a predetermined range in the suction line to thecompressor 10. The sequence of temperature changes in this mode is shown graphically inFIG. 4 , wherein the hot gas temperature is successively increased from the level (A) provided by the full open proportional valve to the higher level (B) at theHEX 86 output and then the final highest level, from theheater 88. - At the opposite (cold) end of the operable temperature range, there are limitations on the low range of temperature possible, depending on the proportion of liquid in the refrigerant mix that is fed to the load. The presence of gas in the saturated mix employed on cooling adversely affects performance by increasing the pressure drop between
load 28 and input tocompressor 10. The temperature of a mix of liquid and vapor at any point is equal to the saturation temperature of the liquid at the pressure experienced by the mix at that particular point. In these systems, with respect to flow in the return line from theload 28 to thecompressor 10 input, the pressure drop is proportional to the square of the mass flow of the refrigerant. Since the cooling output power of a vapor cycle system is proportional to the compressor input pressure, it is advantageous to reduce the mass flow return to the compressor. Specifically, cutting the mass flow in half reduces the pressure drop four-fold, since pressure drop in a flowing gas is about proportional to the square of the mass of the gas flow. - With these factors in mind, the condensed
refrigerant line 16 includes, subsequent to theTXV 24, acheck valve 98 and avapor separator 100, an example of which is seen in more detail inFIG. 2 . Vapor which is collected in theseparator 100 is directed along a vapor line to thereturn line 30 to thecompressor 10. For maximum cooling effect, the liquid proportion is accentuated by thevapor separator 100, and increases the efficiency of thecompressor 10. -
FIG. 2 shows a schematic cross-sectional diagram of theliquid separator 100. Somewhat similar devices are commonly used in vapor-cycle refrigeration systems to separate high pressure refrigerant gas emerging from the compressor from oil mist carried along with the refrigerant but this use in a thermodynamic function is novel. Gas and liquid enter theseparator body 150 atentry port 153 as indicated byarrow 154 onFIG. 2 . Theseparator 100 functions by using a finely divided metal wool barrier 156 placed in the path of the gas such that all the refrigerant must pass through the wool 156 toward an outlet. Droplets coalesce on the surfaces of the metal wool barrier 156 and descend, under the influence of gravity, to the bottom of the separator cavity. When a sufficient level of liquid builds up in the bottom it lifts afloat 160 that is connected to avalve 162 fitted into an exit port 164 and thus allows liquid to flow from the system through the liquid exit port to theload 28. The gaseous refrigerant that passes the barrier 156 flows toward thecompressor 10 input from agas exit port 166 near the top of the separator cavity. Sufficient pressure drop from the inside of theseparator 150 to the outside must be maintained in order to drive the fluids through their respective openings. This is provided for by the use of aninput orifice 153, whose impedance is chosen to be high enough to provide the needed level of driving pressure across the liquid port 164 and gas orvapor port 166. - When liquid only is fed to the mixing
tee 22 to combine with hot gas regulated in flow byproportional valve 22 the combined mixture flowing from mixingtee 26 will simply have less gas than if theseparator 100 were not present. When the maximum amount of cooling is demanded at the lowest possible load temperature a condition is encountered that should be noted.Proportional valve 22 would be shut in this mode and the flow through the system shown inFIG. 1 would be almost pure liquid entering and leaving mixingtee 26. Under some conditions, such as when the refrigerant condenses at 60° C. and provides cooling at −20° C., without theseparator 100 there would only be about 40% of the total mass flow traveling out of mixingtee 22. This means that the total pressure drop over the loop from the mixingtee 26 to the exit line will be only about 16% of what it would be if the full mass flow of gas and liquid were to be passed through the system. This can provide a significant improvement in system performance. A typical pressure drop of such a system, measured from the supply line to the exit line, would be of 12 psig (measured cooling 5 KW at a set point of −40° C.) but would be less than 4 psig with the improved system shown inFIG. 1 . Temperature measured at load would therefore drop about 12° C. when tested under these same conditions. - The basic TDSF system enables providing useful heat to maintain the load in the range of 90° C. to 120° C. by delivering high pressure gas to the load at temperatures that are sometimes well in excess of the required load temperature level. Thus, for example, to provide 5 KW of heat to a load which is to be heated to 120° C. with a flow of 200 grams/second of R507 gas requires that the gas be heated to about 28.5° C. or more above 120° C. In giving up heat to the load by cooling 28.5° C. this gas flow will bring the load to the target temperature. Somewhat more than this amount is needed to provide drive for the needed transfer of heat across whatever heat exchanger is used. A temperature of 120° C. is as high as a typical commercial compressor readily withstands, whereas the improvement of
FIG. 1 can provide gas at temperatures approaching 200° C. if the structural members used can support such levels. - Basically, this is achieved by using the
counter-current HEX 86 together with the extraelectrical heater 88 in the input path to theload 28, after themixer 26. During operation, when the system is supplying temperatures less than about 60° C. the system functions substantially the same as does the prior system. When temperatures above this level are required the hot gas from thecompressor 10 first provides its maximum level [(A) inFIG. 4 .] which is raised to a higher level by flowing through the countercurrent HEX [(B) inFIG. 4 ]. Finally, theheater 88 is activated byelectronic controller 20 to provide adequate heat to raise the temperature of the refrigerant to the desired final value [(C) inFIG. 5 ]. Thecounter-current HEX 86 isolates the bulk of the TDSF system from any adverse effects of the high temperature because the fluid emerging on the return line from thecounter-current HEX 86 will be not much hotter than the fluid emerging from the high pressure outlet of the compressor. - The system includes a further improvement providing adequate control during times when the TDSF system is closely controlling the temperature of an object that is being temperature controlled by the TDSF. A full flow of gas from
proportional valve 22 overwhelms the controller function if the entire flow is mixed with the flow fromTXV 24. This condition is illustrated by the graph ofFIG. 3 , which shows the instability that exists when there is full hot gas flow. The effect of a small change in flow fromvalve 22 is then such as to change the total thermal output of the mixture to an excessive degree, and the system can tend to become unstable. In essence, the loop gain of the servo system which includes the combined output of thevalves valve 24. This causes problems when the temperature of the load is being controlled to close tolerances: A swing of temperature around the control point results, particularly when the controlled load is located at a distance from the point of application of cooling or heating. This condition introduces a time delay between the application of cooling at theload 28 and the reaction of any temperature sensor located at theload 28. - The solution used is to employ a
bypass line 103 including a hot gas bypass (solenoid)valve 104 as shown inFIG. 1 . The hotgas bypass valve 104 is responsive to thecontroller 20. When close control is needed, thevalve 104 is opened, allowing some of the gas output ofproportional valve 22 to bypass directly to the input ofcompressor 10, which has the effect of reducing flow through the cooledload 28. Acheck valve 106 in the bypass line prevents any flow from theTXV 24 from being bypassed when theproportional valve 22 is closed. Thus the pressure drop through theload 28 is reduced and concomitantly the temperature difference across the load is also reduced. Control is enhanced because the overall loop gain of the control servo circuit is reduced and thus easier to control. - These expedients all contribute in a highly integrated fashion to assuring greater reliability and extended range for TDSF systems. Practical applications of this concept can use the potential for fast response and precise control afforded by the system to achieve superior results for particular situations. Some testing and instrumentation systems, for example, test a multiplicity of parts or products sequentially at a series of different temperatures, which may vary widely. The capability of a TDSF system for changing rapidly between temperature levels can save much time and money and increase throughout in these inspection applications.
- It has been found that TDSF systems can respond to needed temperature changes even faster than the electronic controllers, when the controllers have to store a series of readings before establishing reaching a steady state condition. In a typical controller using proportional and derivative functions, for example, the entry of a new set point can initiate a time consuming sequence in which, while transitioning to a new target value, a succession of readings are required. Where a TDSF system has a faster response it has been found useful to enter an artificial and temporary temperature reading into the controller. A new sequence of readings is not needed because previously taken temperature measurements are retained and the controller operates without interrupting the prior sequence. This enables final temperature adjustment of the saturated fluid much more rapidly. In a specific example, the artificial temperature input is used to compensate for thermal delays that are inherent in the design of a tool. For the semiconductor application, there is a physical distance between the top of a chuck, on which the semiconductor wafer rests, and base region where thermal transfer with the refrigerant takes place. By altering the input temperature artificially in step-wise fashion before starting application of power, control of the chuck temperature is both more rapid and precise. Other empirically derived artificial inputs may be used in other situations, for start-up or shut-down sequences.
- While a number of forms and alternatives have been described above, it will be appreciated that the invention is not limited thereto but includes all variants and alternatives within the scope of the appended claims.
Claims (15)
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US11/591,465 US20070095097A1 (en) | 2005-11-03 | 2006-11-02 | Thermal control system and method |
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