Abstract:
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.

Description:
REFERENCE TO PRIOR APPLICATIONS  
       [0001]     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”. 
     
    
     BACKGROUND OF THE INVENTION  
       [0002]     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.  
         [0003]     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.  
         [0004]     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.  
       SUMMARY OF THE INVENTION  
       [0005]     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.  
         [0006]     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.  
         [0007]     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.  
         [0008]     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.  
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0009]     A better understanding of the invention may be had by reference to the following description taken in conjunction with the accompanying drawings, in which:  
         [0010]      FIG. 1  is a block diagram representation of a temperature control system using direct transfer of saturated fluids in accordance with the invention;  
         [0011]      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 ;  
         [0012]      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  
         [0013]      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 . 
     
    
     DETAILED DESCRIPTION OF THE INVENTION  
       [0014]     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.  
         [0015]     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.  
         [0016]     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.  
         [0017]     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.  
         [0018]     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.  
         [0019]     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.  
         [0020]     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 .  
         [0021]     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 .  
         [0022]     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.  
         [0023]     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 .  
         [0024]     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.  
         [0025]     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 .  
         [0026]      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 .  
         [0027]     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.  
         [0028]     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.  
         [0029]     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.  
         [0030]     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 .  
         [0031]     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.  
         [0032]     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.  
         [0033]     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.  
         [0034]     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.