US3264841A

US3264841A - Refrigeration system with means to prevent compressor surge

Info

Publication number: US3264841A
Application number: US377316A
Authority: US
Inventors: William T Osborne
Original assignee: Carrier Corp
Current assignee: Carrier Corp
Priority date: 1964-06-23
Filing date: 1964-06-23
Publication date: 1966-08-09
Anticipated expiration: 1983-08-09

Description

9, 1966 w. "r. osBoRNE 3,264,843

REFRIGERATION SYSTEM WITH MEANS TO PREVENT COMPRESSOR SURGE Filed June 23, 1964 2 Sheets-Sheet 1 INVENTOR. WILLIAM T. OSBORNE.

"WWWW' ATTORNEY.

Aug. 9, 1966 W. QSBORNE 3,264,341

REFRIGERATION SYSTEM WITH MEANS TO PREVENT COMPRESSOR SURGE Filed June 23, 1964 2 Sheets-Sheet 2 FULLY OPEN MODULATING FULLY CLOSED ENTERING conosusms wATER TEMPERATURE SURGE CHARACTERISTICT DEGREES F. STEAM CONDENSATE TEMPERATURE DEGREES F. 10-

0 20 40 so 80 I00 I20 I. NOMINAL COOLING CAPACITY FIG. 2

INVENTOR. WILLIAM T. OSBORNE.

ATTORNEY.

United States Patent 3,264,841 REFRIGERATION SYSTEM WITH MEANS T1) PREVENT COMPRESSOR SURGE William T. Osborne, Syracuse, N.Y., assignor to Carrier Corporation, Syracuse, N.Y., a corporation of Delaware Filed June 23, 1964, Ser. No. 377,316 8 Claims. (Cl. 62-196) This invention relates to means for effectively preventing compressor surge in a refrigeration system and, more particularly, to a hot gas bypass control for a variable speed centrifugal refrigerant compressor.

Various expedients are known in the refrigeration field for preventing compressor surge in a refrigeration system. For example, a hot gas bypass having a normally closed valve may be provided between the high and low pressure sides of the system.

In a refrigeration system disclosed in a copending patent application of Louis H, Leonard for a heating and cooling system, application No. 377, 258, and filed on the same date as the present application, neither the temperature nor flow of condensing water need be controlled. However, as the temperature of condensing water changes, the surge point characteristic of the refrigerant compressor changes. For example, should the condensing water temperature rise, the refrigerant pressure on the high side of the system will rise, and may eventually cause the compressor to become unstable, or to surge, with little or no decrease in compressor speed or refrigerant flow. Should the condensing Water temperature drop, the speed or flow may be decreased a much greater amount before surge occurs. The cooling capacity of the system described in the abovementioned copending application is regulated by regulating the discharge pressure of a steam turbine which drives the compressor to vary the refrigerant output of the compressor. Steam from the turbine discharge passes into a steam condenser, and the turbine discharge steam pressure is regulated by controlled blanketing of a condensing portion of the steam condenser with a noncondensible vapor, preferably refrigerant vapor. For example, when it is desired to reduce the cooling capacity of the system, the output of the turbine and compressor is reduced by increasing the blanketing of the steam condensing portion, thus reducing its condensing capacity and raising the turbine discharge pressure to reduce the turbine and compressor output.

The present invention is directed to etfectively prevent compressor surge in such a system responsive to a steam condensate temperature range.

It is a primary object of this invention to provide a new and improved refrigeration system. A related object is provision in such a system of a new and improved surge prevention control for a refrigerant compressor and, more particularly, for a compressor in a refrigeration system as described in the aforementioned copending patent application. A related object is provision of a surge prevention control which is responsive to steam condensate temperature.

A more specific object is to provide a new and improved surge prevention control for a refrigerant compressor in a refrigeration system wherein a refrigerant side of the system includes a hot gas bypass operable for effectively preventing compressor surge, and a power side of the system includes a steam turbine for driving the compressor and discharging steam into a steam condenser, the output of the turbine and compressor being regulated inversely of the cooling capacity of the system and the steam condensate from the turbine being at a substantially predetermined temperature as the compressor approaches surge condition, and a sensor operable responsive to the steam condensate approaching the predetermined temperature to operate the hot gas bypass for effectively pre venting surge of the compressor.

Additional objects and advantages of the invention will be apparent from the following description and drawings in which:

FIGURE 1 is a flow diagram of a refrigeration system incorporating a preferred embodiment of a compressor surge prevention control;

FIGURE 2 is a graph illustrating compressor surge and surge control characteristics in the system.

The illustrated refrigeration system is preferably air tight and may be considered as having a power side including a circuit for the circulation of power fluid, and a refrigerant side including a circuit for the flow of refrigerant under the influence of drive means driven by the power fluid, with the operation of the system regulated by a control system.

The invention will be described with reference to a preferred power fluid which is water, and a preferred refrigerant which is octafluorocyclobutane, commonly referred to as C318 and having a chemical formula C 1 These fluids are particularly preferred because of their relative immiscibility and because they are inherently highly stable and do not tend to decompose or chemically react with each other or other materials in the system, or cause or promote corrosion and undesirable byproducts. Also, this refrigerant is a relatively noncondensible vapor at the temperatures and pressures at which the power fluid (water) condenses as Well as at the usual ambient atmospheric conditions of temperature and pressure. However, other power fluids and refrigerants having the desired chemical and physical properties may be utilized within the scope of this invention.

With reference to FIGURE 1, the power side of the refrigeration system includes a suitable steam generator 11 Which supplies steam through a steam supply line 12 to steam powered drive means here in the form of a turbine section 13 of a turbocompressor 13. The turbine 13 discharges saturated steam into a discharge steam line 15 opening into a steam condenser 16. The steam is condensed by a first condensing portion or tube bundle 17, and a second condensing portion or tube bundle 18 which provides heat to a load to be heated. A heating water pump 18' recirculates heated water from the second condensing portion to the load. Steam condensate in the condenser 16 passes through a port 19 and into a steam condensate chamber 20. From the chamber 20 the steam condensate is returned through a steam condensate line 21, by means of a steam condensate pump 22 in the line, to the steam generator 11 for recirculation through the power side of the system. In the illustrated embodiment, the steam is supplied to the turbine at a substantially constant pressure, for example 15 p.s.i.g., as by means of a suitable constant pressure steam regulating valve 23 in the steam supply line 12.

The turbocompressor 13' is preferably provided with water lubricated bearings. The steam condensate pump 22 passes steam condensate through a lubricant line 24 and cooling means 24' to the turbocompressor, from which the lubricating water and any leakage of refrigerant or steam within the turbocompressor is returned to the steam condensate chamber 20 through a return line 25.

The refrigerant side of the system includes a refrigerant compressor section 31 of the turbocompressor 13', and preferably a centrifugal compressor for discharging relatively high pressure refrigerant vapor from the compressor outlet 32 into a refrigerant line 33 opening into a refrigerant condenser 34. The refrigerant condenser 34 has a condensing portion or tube bundle 35 for circulating condensing water to condense the refrigerant vapor. Refrigerant condensate passes from the refrigerant condenser through a refrigerant condensate line 36 to a refrigerant subcooler 37 and then through a line 39 to suitable refrigerant flow control means 40, here in the form of a float valve assembly. In keeping with normal practice, an equalizer line 40 may be provided between the refrigerant condenser 34 and the chamber of the flow control means 30. The refrigerant then passes through a cooler supply line 41 and into a cooler 42, and more particularly, into a refrigerant pan 43. The pan 43 contains a flooded chilled water tube bundle 44 communicating with a chilled water line 45 having a chilled water pump 46 for circulating chilled water to a load to be cooled. Boiling refrigerant in the pan vaporizes into a refrigerant chamber 47 of the cooler and is withdrawn through a suction line 49 to an inlet connection 49 of the compressor 31. The portion of the refrigerant side of the system between the compressor outlet 32 and the float valve 40 defines a high pressure side, and the portion between the float valve 40 and the compressor inlet 49' defines a low pressure side.

The refrigerant side of the system further includes a hot gas bypass operable for effectively preventing compressor surge and including a hot gas bypass line 55 cnmeeting the cooler supply line 41 and an upper portion of the refrigerant condenser 34 for passing refrigerant vapor around the float valve 40 when a normally closed modulating refrigerant valve 56 in the bypass line 55 is open.

Cooling capacity control of the system is regulated by regulating the refrigerant output of the compressor 31 which is connected with the turbine 13 by a shaft 60. In order to regulate the power output of the turbine 13, the pressure within the steam condenser 16 is regulated. Refrigerant, which is a noncondensible vapor within the steam condenser 16 is passed through a refrigerant line 62 from the refrigerant chamber 47 of the cooler 42 and through an inlet port 63 in the steam condenser to blanket the first condensing bundle 17 of the steam condenser with refrigerant vapor and reduce the condensing capacity, thereby raising the condenser pressure to reduce the turbine output, as is more fully described in the previously mentioned Leonard application.

A baffie 64 between the tube bundles 17 and 18 extends longitudinally through the steam condenser in sealed engagement with a steam condenser shell 64' except at a limited area of communication 65 at the refrigerant port 63 and at an end of the bundles opposite the outlet port 19 and the opening of the steam discharge line into the upper portion of the steam condenser, so that the second bundle 18 is free of refrigerant vapor to provide maximum possible heat output, while the first bundle 17 is blanketed by refrigerant vapor. Steam condensate passes through the outlet port 19 and into the steam condensate chamber from which the condensate is returned to the steam generator 11.

Refrigerant vapor is withdrawn from the steam condenser 16 by means of a purge line 70 opening into the steam condensate chamber 20. An opposite end of the purge line 70 opens into the throat of a jet pump 71 in a lower water sump portion 72 below the pan 43 within the cooler 42. Water entering the cooler 42 is separated from refrigerant and collects in the sump 72, as is more fully described in the previously mentioned Leonard application and in my copending United States patent application for a cooler, application No. 377,317, and filed on the same date as the present application. Water is recirculated through the sump by a suitable water supply pump 73 which provides impeller water for the jet pump 71. In the preferred embodiment, the water supply pump 73 operates at constant speed so that the purge line 70 withdraws refrigerant vapor from the steam condenser at a substantially constant rate. A modulating refrigerant valve 75 in the refrigerant line 62 is preferably controlled by a suitable temperature sensor 76 on the chilled water line 45 so that the quantity of refrigerant entering the steam condenser, and therefore the turbine and compressor outputs, are controlled responsive to chilled water temperature.

Make-up water for the power side may be provided by a make-up water line 77 from the water supply pump 73 to the steam condensate chamber 20. A float sensor 78 in the chamber 20 operates a valve 79 in the line 77 to add make-up water as needed.

A suitable condensing medium, such as tower water, is circulated by a tower water pump 80 from a supply line 81 connected with the tower (not shown) and through the refrigerant subcooler 37 and an intermediate line 82 to the refrigerant condenser tube bundle 35 and then the steam condenser first tube bundle 17, and back to the tower through a return line 83. Thus, the tower water is serially circulated through the refrigerant condenser and then the steam condenser.

Neither tower water temperature nor flow rate need be regulated during operation of the system. However, as the tower water temperature rises, it can effect less condensing and compressor surge may occur. The refrigerant pressure in the refrigerant condenser 34 rises, thus increasing the compressor discharge pressure. Also, the steam condenser pressure increases thus increasing the turbine discharge pressure and the temperature of the steam entering the steam condenser to increase the steam condensate temperature.

Liquid subcooling is a characteristic of condensers containing noncondensibles and is defined as the difference in temperature between the liquid leaving the condenser and the actual condensing temperature inside the condenser. Subcooling occurs by virtue of the inability of the saturated vapors to reach the colder tubes which are blanketed by noncondensibles and the tremendous capability of the blanketed, colder tubes to cool the droplets of liquid condensate as they drop, by gravity, from higher non-blanketed tubes, or as the droplets are carried into the blanketed tubes by the motion of the vapors and gases in the condenser. Generally, the amount of subcooling which occurs is proportional to the quantity of noncondensibles present and effectively blanketing the condensing tubes, and to the difference between the actual condensing temperature and the temperature of the water passing through the condenser tubes. The steam condensate temperature is a function of the amount of subcooling and the actual condensing temperature. However, as the refrigerant condenser pressure increases, the compressor 31 must pass a greater quantity of refrigerant through the system at a greater pressure differential in order to maintain the same cooling capacity. Thus, the output of the compressor 31 must be increased, resulting in the refrigerant condenser pressure increasing and requiring a further increase in compressor output. In order to increase the compressor output, the turbine discharge pressure must be reduced by reducing the blanketing of the first condensing bundle 17, resulting in a reduction of the partial pressure of the noncondensible refrigerant vapor and a reduction of condensate subcooling. The end result is that the temperature of the steam condensate leaving the steam condenser 16 rises appreciably as the compressor approaches its surge point.

Another condition which may result in compressor surge is low cooling capacity at which time a relatively small quantity of refrigerant vapor is passed by the compressor '31 so that the compressor speed is necessarily slow. Thus, the turbine must operate at slow speed necessitating a high turbine discharge pressure resulting from very great blanketing of the steam condenser first condensing bundle 17, thus producing a high steam condenser pressure w-hich results in high turbine steam exhaust temperature and a relatively high steam condensate temperature.

Therefore, in order to retard the tendency of the compressor 31 to surge under such conditions, the temperature of steam condensate leaving the steam condenser 16 may be sensed for opening and regulating the amount of opening of the hot gas bypass valve 56. The temperature of steam condensate leaving the steam condenser is sensed by a temperature sensor 85 in the steam condensate chamber 20. The sensor 85 is preferably a selfcontained portion of the hot gas bypass valve 56.

With reference to FIGURE 2, the surge characteristic of the compressor 31 and a family of steam condensate temperature curves is plotted against entering tower or condensing water temperature and percentage of nominal cooling capacity. If the turbocompressor is operated without suitable surge prevention means, the compressor will surge at any combined condition of tower water temperature and percentage cooling capacity to the left of the surge characteristic curve. The family of steam condensate temperature curves closely parallel the surge characteristic curve, and the sensor 85 in the steam condensate chamber 20 is adjusted to open the hot gas bypass valve 56 when the steam condensate temperature in the chamber reaches 122 F, for example. At a steam condensate temperature of 126 F., for example, the hot gas bypass valve 56 is fully open and modulates between the closed and full open positions in proportion to the temperature of the steam condensate between 122 F. and 126 F. Because the condensate temperature and surge characteristic curves closely parallel each other,.

greater operating efficiency of the system throughout a greater range of operating conditions may he obtained by means of the subject surge prevention control.

The following chart indicates various operating conditions throughout the system:

Nominal Cooling Capacity 100% 50% Entering Condensing Water,

Lv. Chilled Water, F 44 43 42 41. 40. 5 40 Cooler:

F 36 35 34 33 32. 5 32 p.s.i.a 19. 7 19. 2 18. 7 18. 2 18 17. 7 Steam Condenser:

105 85 95 75 86 66 p.s.i.a 70 51 60 43 52 36 Ref. Leaving Subcooler, F 95 75 90 70 86 66 In summary, during operation of the refrigeration system with relatively high tower water temperatures, a reduction of required cooling capacity is sensed by the control valve sensor 76, which regulates the rate of bleeding noncondensibles into the steam condenser to raise the turbine discharge pressure and reduce the turbine power output. As the cooling load continues to decrease, and the steam condenser pressure is correspondingly increased, the compressor approaches a surge condition. At the same time, the steam condensate temperature, which is a function of steam flow, unblanketed steam condensing surface, and tower water temperature in the steam condenser, rises to a relatively high determinable level. During conditions of decreasing capacity with relatively low tower water temperature, the steam condenser pressure is increased more for each reduction of capacity since less turbine power is required at each capacity condition. However, since the steam condensate temperature is influenced by the temperature of the tower water in the steam condensing bundle, it increases less rapidly with capacity reduction at low tower water temperatures. When the compressor reaches surge condition with lower entering tower water temperature, the steam condenser pressure is higher, the cooling capacity is less, and the steam condensate temperature is the same as with high tower water temperatures.

It should be noted that in the foregoing system, the lower the tower water temperature the lower may be the partial 'load operation of the compressor without entering the surge range. With relatively cold tower water, much lower cooling capacity of the system is necessary to produce a sufficiently high steam condensate temperature for operating the hot gas bypass valve. Should the temperature of the tower water increase, the condensing capacity of the steam condenser will be reduced so that less blanketing of the steam condenser is required to maintain a desired turbine discharge pressure for a particular cooling capacity of the system, and while the temperature of the discharged steam remains the same, the higher temperature of the steam condensing portion tends to produce a higher steam condensate temperature. Thus, as compressor surge condition is reached the steam condensate approaches a relatively high determinable temperature.

While a preferred embodiment of the invention has been described and illustrated it will be understood that the invention is not limited thereto since it may be otherwise embodied within the scope of the following claims.

Iclaim:

1. In a refrigeration system, the combination comprising, a refrigerant compressor, first means operable for effectively preventing compressor surge, a steam condenser, steam powered drive means for driving said compressor and discharging steam into said steam condenser, second means for regulating the output of said drive means and thereby the output of said compressor in proportion to the cooling capacity of the system and as said output is reduced, for raising the steam condensate to a substantially predetermined temperature as the compressor approaches surge condition, and control means responsive to said predetermined condensate temperature for operating said first means and efiiectively preventing surge of said compressor.

2. The system of claim 1 wherein the output of said compressor is proportional to the output of said drive means, the drive means output being variable inversely of its discharge pressure, and said second means including, means for passing a nonoondensible vapor into said steam condenser to regulate the steam condenser pressure and steam condensate temperature, and means for regulating the quantity of said nonoondensible vapor in said steam condenser inversely of the cooling capacity of the system.

3. The system of claim 2, and a refrigerant condenser on a high side of the system and a cooler on \a low side of the system, the refrigerant and steam condensers having condensing portions for circulation of a condensing medium, whereby at a relatively high entering condensing medium temperature the refrigerant condenser pressure is relatively high tending to cause compressor surge and the steam condensate temperature is relatively high as the compressor approaches surge condition.

4. The system of claim 3 wherein said condensing portions are in series, whereby said condensing medium is serially circulated through said condensing portions.

5. The system of claim 4, and means for circulating said condensing medium through the refrigerant condensing portion first.

6. The system of claim 5, and said first means comprising hot gas bypass means between said high side and said low side.

7. The system of claim 3 wherein said control means operates said hot gas bypass means for passing refrigerant at a flow rate in proportion to the steam condensate temperature above said predetermined temperature.

8. The system of claim 7 wherein said compressor is a centrifugal compressor.

References Cited by the Examiner UNITED STATES PATENTS 2,183,821 12/1939 Nelson 62500 X 2,983,111 5/1961 Miner 62-228 X MEYER PERLIN, Primary Examiner.

Claims

1. IN A REFRIGERATION SYSTEM, THE COMBINATION COMPRISING, A REFRIGERANT COMPRESSOR, FIRST MEANS OPERABLE FOR EFFECTIVELY PREVENTING COMPRESSOR SURGE, A STEAM CONDENSER, STREAM DRIVE MEANS FOR DRIVING SAID COMPRESSOR AND DISCHARGING STEAM INTO SAID STEAM CONDENSER, SECOND MEANS FOR REGULATING THE OUTPUT OF SAID DRIVE MEANS AND THEREBY THE OUTPUT OF SAID COMPRESSOR IN PROPORTION TO THE COOLING CAPACITY OF THE SYSTEM AND AS SAID OUTPUT IS REDUCED, FOR RAISING THE STEAM CONDENSATE TO A SUBSTANTIALLY PREDETERMINED TEMPERATURE AS THE COMPRESSOR APPROACHES SURGE CONDITION, AND CONTROL MEANS RESPONSIVE TO SAID PREDETERMINED CONDENSATE TEMPERATURE FOR OPERATING SAID FIRST MEANS AND EFFECTIVE PREVENTING SURGE OF SAID COMPRESSOR.