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Publication numberUS20070095097 A1
Publication typeApplication
Application numberUS 11/591,465
Publication dateMay 3, 2007
Filing dateNov 2, 2006
Priority dateNov 3, 2005
Publication number11591465, 591465, US 2007/0095097 A1, US 2007/095097 A1, US 20070095097 A1, US 20070095097A1, US 2007095097 A1, US 2007095097A1, US-A1-20070095097, US-A1-2007095097, US2007/0095097A1, US2007/095097A1, US20070095097 A1, US20070095097A1, US2007095097 A1, US2007095097A1
InventorsKenneth Cowans, William Cowans, Glenn Zubillaga
Original AssigneeCowans Kenneth W, Cowans William W, Zubillaga Glenn W
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Thermal control system and method
US 20070095097 A1
Improvements in systems and methods which employ a two-phase fluid, such as a refrigerant, as a saturable fluid in direct heat transfer relation to a thermal load, are realized by extraction of vapor from the saturated fluid before heat exchange. Moreover, automatic changing of the paths under command of a controller enables charges to be effected between different modes at higher rates than in other systems by employing the variety of modes available in the direct transfer system.
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1. A method for the efficient exchange of thermal energy between a source and a load in which a two-phase medium comprising a saturable fluid is utilized that has a controllable temperature, vapor and pressure state comprising the steps of:
providing a pressurized saturable fluid at a temperature that is pressure determined;
before delivery of the saturable fluid to the load, extracting vapor from the saturable fluid while maintaining the pressure;
exchanging thermal energy between the saturable fluid after vapor extraction and the load;
recombining the extracted vapor with the fluid after thermal energy exchange and,
restoring the fluid to a selected pressure and temperature status.
2. The method as set forth in claim 1 further including the steps of:
compressing the fluid to a high pressure state;
condensing a portion of the high pressure fluid to substantially ambient temperature;
expanding the condensed fluid to a controlled pressure-temperature state;
blocking return flow of extracted vapor into the expansion step;
mixing the expanded fluid with compressed uncondensed fluid in selected proportions to provide a final temperature state for thermal energy exchange with the load;
exchanging thermal energy between the mixture and the load thereafter;
recombining the extracted vapor with the fluid after thermal energy exchange with the load, and
recompressing the recombined fluid.
3. The method as set forth in claim 2 above, wherein the extraction of vapor reduces the mass flow of saturable fluid through the load, with more efficient thermal exchange, and the saturable fluid after extraction is more efficiently recompressed.
4. The method as set forth in claim 2 above, further including the steps of controlling the proportion of compressed uncondensed fluid relative to the expanded fluid to determine the temperature thereof, sensing the change of load temperature when full hot gas flow is employed, and stabilizing the load temperature by diverting a partial flow of compressed uncondensed fluid directly for recompression.
5. The method of circulating a two-phase fluid through a thermal exchange procedure so as to maintain efficiency over a wide temperature range of operation comprising the steps of:
compressing the fluid to a high pressure hot gas state;
condensing a portion of the pressurized hot gas to a liquid;
selectively expanding at least some of the condensed liquid to a liquid/vapor phase;
extracting vapor from the liquid/vapor phase while maintaining the pressure thereof;
mixing a selected proportion of the pressurized hot gas with the expanded liquid/vapor phase after extraction of vapor to establish a chosen temperature/pressure status in the liquid/vapor mixture prior to the thermal load;
passing the mixture in thermal exchange relation with the load, and
recycling the fluid from the load for recompression.
6. A method as set forth in claim 5 above, further including the steps of accumulating refrigerant after thermal exchange with the thermal load, and heating the accumulated refrigerant to establish desired pressure and temperature balance for recompression, and diverting a flow of pressurized hot gas directly to recompression and redirecting the flow passed in thermal exchange with the load.
7. A method as set forth in claim 5 above, wherein the step of condensation comprises supplying a condensate in thermal exchange with the compressed fluid, sensing the temperature of the condensate after the thermal exchange and adjusting the flow rate of coolant to maintain the temperature of the condensate after thermal exchange at a selected level, and wherein the step of extracting vapor includes coalescing droplets of vapor, accumulating a reservoir of liquid refrigerant therefrom, and flowing refrigerant in excess of a given level to thermal exchange with the load.
8. A method for improving the control stability of a system utilizing saturable fluid in compressed, controlled temperature phase in a vapor-compression cycle, for thermal exchange with a load, comprising the steps of:
compressing a saturable fluid to a pressurized hot gas;
condensing a flow of the pressurized hot gas;
expanding the condensed flow to a selected level;
reducing the vapor content of the expanded flow without changing the pressure;
mixing a selected proportion of pressurized hot gas with the expanded flow;
adding thermal energy to the mixture to reach a target temperature;
feeding the mixture to the load for thermal exchange prior to return for recompression, and
when the proportion of hot gas to be provided to the load is at a high level, shunting a selected fraction therefrom directly to compression to stabilize the functions of controlling the expansion and controlling the pressure.
9. A device for isothermally separating vapor from a liquid/vapor fluid mixture at saturation to increase the proportion of liquid therein comprising;
a chamber having input and output ports for the saturable fluid;
a barrier section of selected porosity and high surface area disposed within the chamber between the input and output ports and intercepting the liquid/vapor mixture;
a liquid outlet at a lower region of the chamber for output of liquid collecting within the chamber, and
a valve including a surface float disposed on the surface of liquid in the chamber and including a closure member in operative relation to the liquid outlet that is disposed to allow liquid in excess of a predetermined level in the chamber to flow out of the chamber.
10. A device as set forth in claim 9 above, wherein the inlet port has a flow impedance selected such that flow occurs from within the chamber through the outlet vapor port and the outlet liquid port, and wherein the inlet vapor port provides the saturable fluid and the barrier is a fibrous structure substantially vertically intercepting the flow path between the input and output ports and further including a check valve coupled in the flow path to the inlet port.
11. a system for controlling the temperature of at least one thermal load unit using a circulating two-phase refrigerant comprising:
a vapor cycle refrigeration system including at least one compressor and condenser series, and employing a two-phase refrigerant as a temperature control medium, and arranged such that the refrigerant is first compressed and then at least a portion thereof is separately liquefied to a state of pressure saturation;
a proportioning device receiving a flow of compressed refrigerant as a hot gas;
an expansion device receiving a flow of saturated liquid;
a gas processor receiving the saturated liquefied state, for extracting at least some vapor therefrom without substantially altering the pressure/temperature equilibrium thereof,
a mixing junction coupled to receive the hot gas flow from the proportioning device and the flow from the expansion device;
a thermal load comprising at least one heat exchange structure receiving the vapor-extracted saturated liquid, and
a controller system coupled to operate the proportioning device and the expansion device to control the temperature of the thermal load.
12. A system as set forth in claim 11 above, wherein the extracted vapor from the gas processor is circulated to the compressor as input and wherein the flow path to the compressor also includes an accumulator receiving output from the load, and a controller-operated coupling from the proportioning device to the flow path to the compressor.
13. A system as set forth in claim 12 above, wherein the gas processor comprises a vapor separator including a level-responsive liquid-enriched flow of saturated fluid to the load, and thereafter vapor to the compressor input, and where a check valve is coupled into the input path to the vapor separator.
14. A system as set forth in claim 13 above, further including controller operated shunt and bypass valves in circuit with the proportioning device to enable abrupt commencement and cessation of hot gas flow and controller operated shunt and bypass valves in circuit with the expansion device to enable abrupt commencement and cessation of flow of condensed fluid and a close on rise valve in circuit with the compressor input and responsive to the input pressure thereat.
15. A system as set forth in claim 13 above, wherein the system includes means to extend the high temperature range of operation comprising:
a countercurrent heat exchanger coupled in circuit with the load and exchanging thermal energy between input and output flow to and from the load and a heater under command of the controller for adding thermal energy to the input to the load.

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 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; and

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.


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, 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. In the separate 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 a mixing 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 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.). 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, a shutoff valve 40 is disposed in series with the proportional valve 22. For virtually immediate or assured full flow of hot gas, the shutoff valve 40 is bypassed by a shunt valve 46 which is in parallel with it. For rapid control of the condensed refrigerant flow, 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. In a configuration known in the prior art, the water flow rate is governed by a flow control system 54, so as to maintain the output pressure from the compressor 10. To facilitate maximum cooling, a bypass 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 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. In accordance with the W. W. Cowans Pat. No. 6,446,446 referenced in the predecessor parent application, 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.

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 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. Also, 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.

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 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. For example, 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. For this purpose, 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. Separately, 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.

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 to compressor 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 the load 28 to the compressor 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 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. For maximum cooling effect, 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. When a sufficient level of liquid builds up in the bottom it lifts a float 160 that is connected to a valve 162 fitted into an exit port 164 and thus allows liquid to flow from the system through the liquid exit port to the load 28. The gaseous refrigerant that passes the barrier 156 flows toward the compressor 10 input from a gas exit port 166 near the top of the separator cavity. Sufficient pressure drop from the inside of the separator 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 an input 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 or vapor port 166.

When liquid only is fed to the mixing tee 22 to combine with hot gas regulated in flow by proportional valve 22 the combined mixture flowing from mixing tee 26 will simply have less gas than if the separator 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 in FIG. 1 would be almost pure liquid entering and leaving mixing tee 26. Under some conditions, such as when the refrigerant condenses at 60 C. and provides cooling at −20 C., without the separator 100 there would only be about 40% of the total mass flow traveling out of mixing tee 22. This means that the total pressure drop over the loop from the mixing tee 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 in FIG. 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 extra electrical heater 88 in the input path to the load 28, after the mixer 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 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]. Finally, 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. In essence, 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. 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 the load 28 and the reaction of any temperature sensor located at the load 28.

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. Thus the pressure drop through the load 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.

Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US7415835 *Oct 12, 2006Aug 26, 2008Advanced Thermal Sciences Corp.Thermal control system and method
US7765820Aug 22, 2008Aug 3, 2010Advanced Thermal Sciences, Corp.Thermal control system and method
US8012304Apr 21, 2006Sep 6, 2011Applied Materials, Inc.Plasma reactor with a multiple zone thermal control feed forward control apparatus
US8157951Apr 21, 2006Apr 17, 2012Applied Materials, Inc.Capacitively coupled plasma reactor having very agile wafer temperature control
US8221580Apr 21, 2006Jul 17, 2012Applied Materials, Inc.Plasma reactor with wafer backside thermal loop, two-phase internal pedestal thermal loop and a control processor governing both loops
US8453468 *Mar 18, 2010Jun 4, 2013Be Aerospace, Inc.System and method for thermal control of different heat loads from a single source of saturated fluids
US8532832 *Sep 14, 2009Sep 10, 2013Be Aerospace, Inc.Method and apparatus for thermal exchange with two-phase media
US20100076611 *Sep 14, 2009Mar 25, 2010Advanced Thermal Sciences, Inc.Method and apparatus for thermal exchange with two-phase media
U.S. Classification62/512, 62/513
International ClassificationF25B43/00, F25B41/00
Cooperative ClassificationF25B2400/0403, F25B2400/01, F25B40/00, F25B41/00, F25B2400/0411
European ClassificationF25B41/00
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