Chillers, Chilled Water Systems Principles and Terminology Chapter Three Chiller Main Components/Compressors/Screw Compressors

Screw Compressors: A rotary-screw compressor uses rotors to compress larger volumes of gaseous refrigerant to a high pressure and temperature. The screw compressor is classified as positive displacement compressor, means that a volume of gas is trapped with an enclosed space whose volume is then reduced by male and female rotors as they rotate. Cool vapor refrigerant enters from the suction port, is forced by the meshing rotors through the threads and exits at the discharge port with high pressure and temperature. Rotary-screw compressor can be sub-classified by the quantity of screws (single, twin and multi). Tri-rotor design allows for shorter rotors and are more efficient than the twin-rotor design, they also balance thrust, which will reduce bearing loads. Twin rotary-screw compressor is designed to be oil flooded machine therefore they are incorporated with oil injection system (system include oil separator, oil pump, oil cooler and oil heater). Most of rotary-screw compressors have a type of capacity control, also they can use an economizer (economizer will increase efficiency and capacity).

Conventional rotary screw compressors are composed of two parallel rotors with external helical profiles fit into a casing. One of the rotors is coupled to the motor (drive rotor or male) and as it turns it moves the other rotor (driven rotor or female), like a common gear set, it is difficult to visualize the geometric profile of the rotating rotors, but we can relate the compression process to a reciprocating compressor, if we consider the drive rotor as the piston and the driven rotors as the cylinder. As the drive rotors and driven rotors unmesh, an empty cylinder is created, drawing in suction gas through the synchronized opening on the rotor suction face. As rotation continues, the suction and discharge rotor faces are sealed off, trapping the gas in the cylinder. When this happens, the meshing point moves toward the discharge end of the rotors and drives the gas ahead of it, see figure 3.34 below. The discharge port provided for the gas escape is relatively small, compared to the suction port.

FIGURE 3.34. Twin screw compressor principle of operation.

Heinreich Krigar of Germany developed the first screw compressor in 1878. In the early 1930’s, a Swedish engineer by the name of Alf Lysholm developed the profile of the modern screw compressor for gas and steam turbine applications. Screw compressors have been used in HVAC applications for nearly five decades. The first compressor designated and install on a chiller in 1967, was made by a US manufacturer called Dunham-Bush under the leadership of a dynamic entrepreneur named Cecil Boling, the first president of ASHRAE, see figure 3.35

FIGURE 3.35. Design of first large hermetic screw compressor.

Rotary screw compressors are well known for their robustness, simplicity, and reliability. They are designed for long periods of continuous operation, needing very little maintenance, they can overcome high lift when speed is reduced, allowing energy savings without the possibility of surge as the compressor unloads. It is used in the commercial and industrial air conditioning and refrigeration applications with a capacity range from 20 to 750 tons (70-2,637 kW).

Screw Compressor Benefits and Limitations

?The rotary-screw compressor has the following benefits compared to other compressors used in chillers:

• Simpler structure, less components, larger capacity, and higher efficiency.
• Less vibration and less surging due to continuous gas displacement via the sweeping motion of the rotors.
• Better adjustment in cooling capacity without causing unstable operation, which is sometimes an issue with centrifugal compressors.
• Less sensitive to liquid slugging but long-term liquid slugging will impact the reliability of the compressor.
• Flatter volumetric efficiency curves as a function of compression ratio (CR) resulting in significant reduction in the required compressor displacement compared to the reciprocating type at high CR typical of medium and low temperature refrigeration, as well as heat pump applications.
• Use of an economizer system by injecting refrigerant vapor into the compressor at a suitable intermediate pressure level resulting in increased capacity and better energy ratio; noting that in the case of screw compressors equipped with a sliding valve unloader or other device based on the same principle the economizer effectiveness quickly vanishes below full load unless the side port is part of the slide valve mechanism as in Bitzer CSH screw compressor design where the port will move with the slid and also using Vi control at part load conditions. The economizer system remains fully effective down to minimum load when capacity control is achieved exclusively by varying the compressor speed with VFD. This feature is also important at higher capacities in competition with centrifugal compressor.

Limitations

Impractical to design to a capacity below 20 tons (70 kW), due to the rotor processing technology.

Screw Compressor Capacity Control and Vi Control

There are two approaches for achieving capacity control with screw compressors technologies one is controlling the suction port and the second is using variable speed motor. Methods for controlling the suction ports are, capacity control slide valve doesn’t control the discharge port, slide valve also control the discharge port, lift valve and slot valve. The single slide valve, double slide valves (for both capacity control and volume ratio control) and variable speed drives are the most common applications.

Slide Valve:

The early design and most common approach for capacity control in screw compressors is the use of a “slide valve” in which it will allow recirculation from the compression chamber back to suction prior to compression (delaying the start of compression internally). Figure 3.36 show the slide valve is placed between the two rotors and consists of two members, one fixed and the other movable. The compressor develops full capacity when the movable portion bears on the fixed member. For capacity reduction the movable portion of the slide separates from the fixed portion so that some of the gas that has filled the cavity during the suction process is not compressed.

FIGURE 3.36. Slide valve for capacity control of a screw compressor: (A) its position relative to the rotors, (B) slide at full-capacity position and (C) slide at reduced-capacity position.

The slide valve permits a smooth, continuous modulation of capacity from full to 10% of full capacity, but that smooth modulation will result in reduction in efficiency at part load.

Variable Volume Ratio Concept

Since screw compressors are volumetric machines with a fixed built-in volume ratio, a mismatch of the built-in volume ratio and that required for the actual compressor pressure ratio results in either over compression or under compression. Mismatch in the built-in volume ratio too large over compression, too small under compression see Figure 3.37, It is rare that the developed pressure within the compressor precisely matches that prevailing in the discharge line. The built-in volume ratio Vi is one of the fundamental characteristics of the basic screw compressor.

Volume ratio=Vi= Vs/Vd

Vi= Built-in volume ratio is the change in volume within the compressor from a larger volume at compressor inlet to a smaller volume at compressor exit.

For a given refrigerant, a fixed volume ratio corresponds thermodynamically to a certain pressure ratio. The higher the volume ratio the larger the corresponding pressure ratio.

Vs= The volume of the gas in cavity when suction port closes

Vd= The volume of the gas in cavity when discharge port uncovers

Pressure ratio= Compression ratio= Pi= Pd/Ps

Pi= Built-in pressure ratio

Pd= Pressure of the gas just before discharge

Ps= Pressure of the gas just before the compression begins

Pi= Built-in pressure ratio

Pd= Pressure of the gas just before discharge

Ps= Pressure of the gas just before the compression begins

FIGURE 3.37. (A) Over-compression Vi too large and (B) Under-compression Vi too small, shown on a pressure-volume diagram.

For a given rotor diameter each different vi is associated with a different rotor length, Some typical values of Vi used by manufacturers are 2.6, 3.6, 4.2, and 5.0. and each Vi corresponds to a certain pressure ratio that varies from one refrigerant to another.

?Since Compressor selections need to take account of the peak operating conditions likely to be encountered, however actual operating conditions may vary, resulting in lower efficiencies. Control of capacity and volume ratio can maintain high efficiency levels.

Built-in volume ratio (Vi) control:

For varying the built-in volume ratio, slide valves or lift valves devices can be used. They can be complicated and expensive, but they will help to avoid the mismatch between Vi and the Pi and will insure an efficient performance over a wide range of operating conditions.

Vi Slide-Valve: A valve, having a sliding action parallel to the rotor bores, situated within the high-pressure cusp region, facing one or both rotor bores and controlling the radial discharge port. The axial port is designed for a value corresponding to the highest Vi required.

Figure 3.38, shows a Vi device function in conjunction with a slide valve that also controls the capacity, consists of two parts which can move independently, one slide valve in the suction region controlling capacity and the other slide valve in the discharge region controlling the Vi. In Figure 3.38-1 the two parts have no gap between them, so no refrigerant vapor vents back to the suction and the compressor operates at full capacity. The discharge port is uncovered when the cavities of the rotors move past the right end of the right member. If the Vi is to be increased but full capacity to be maintained, both parts move to the right, as in Figure 3.38-2, at this position the discharge is delayed so that the pressure in the cavities builds up more before discharging. If the high value of Vi is to be maintained, but the capacity needs to be reduced, the left member backs off which vents some vapor back to the suction, as in Figure 3.38-3.

?The operation of this device requires a complex control, and there are limitations in achieving the desired Vi when the capacity must also be reduced. If the capacity has been reduced by as much as 50%, the variable Vi portion of the control may no longer be able to meet its requirements.

FIGURE 3.38. A variable capacity and Vi device at the following operating conditions: (1) full load and low Vi, (2) full load and high Vi and (3) part load and high Vi.

Vi Lift Valve: A valve within the discharge region, facing one or both rotor bores (including outlet endplate) and opening an extra radial or axial port when a lower Vi is required. The fixed discharge port is determined with regard to the highest-pressure ratio in operation. One Vi lift valve means two-step Vi control, two Vi valves mean three step Vi control, etc. Some of the most efficient available steps, (2.2, 3.5, or 5.0 volume ratio).

Features of the variable built-in volume ratio in screw compressors:

• No oil foam in oil (no over-compression) and therefore less oil carried over into the refrigeration system.
• No overloaded motors or shafts (assuming constant discharge pressure).
• Extended bearing life (minimized load on bearings).
• More efficient liquid refrigerant injection possibilities.
• Extended efficient operating range with economizer (discharge port corrected for flash gas from economizer as well as gas coming from suction).

Capacity Control Using a Variable Speed Motor:

If reducing the capacity using a slide valve result in a decline of compression efficiency, it is reasonable to explore other means of capacity control. The most attractive alternative is the use of variable-speed drive. A positive displacement compressors speed is independent of lift, that mean compressor can develop the same amount of lift at any speed. Therefore, mechanical loaders can be replaced entirely by speed control.

?A variable-speed drive will be provided either by a two-speed motor or by a frequency inverter (see Figure 3.39 for VFD principle of operation) that furnishes infinite variations of speed.

FIGURE 3.39 Frequency inverter principle of operation.

?Motor speed (rpm) is dependent upon frequency. Varying the frequency output of the VFD controls motor speed:

Speed (rpm) = frequency (hertz) x 120 / no. of poles

The draw back when using a VSD at low speeds the ratio of leakage gas to that pumped increases, so both the volumetric and the compression efficiency decrease. At high speeds the high pressure drop through the passages of the compressor reduces the compression efficiency. Therefore, some manufacturers operate the compressors with a combination of slide valve and VFD control. The compressor can be operated at minimum speed of 50% and the slide valve used for further unloading if appropriate discharge and motor temperature limits are observed. The compressor slide valve should remain in the fully loaded position while the VFD is controlling compressor capacity between 50% to 100%. The slide valve may be used to reduce compressor capacity below 50% if required.

From an overall efficiency standpoint, it would be desirable way, to use variable speed screw compressor combined with variable Vi. In this the inherent adaptability of the compressor to variable more speed could mate very well with some degree of variable Vi for efficient variable capacity operation at various compression ratios.

The speed boundaries for the compressor must operate is generally between 5000 to 800 rpm that is a modulation between 100% to about 16%. Additional benefit of a variable frequency inverters is that a frequency higher than the power-line frequency of 60 Hz or 50 Hz can also be developed. This capability permits the handling of peak loads with a compressor slightly smaller than would normally have been necessary.

?Applications where high lift requirements remain even as cooling loads decrease may favor variable speed screw chillers more significantly. An example of this application is a building located in a hot and humid climate where cooling loads can vary while outdoor air temperatures remain high. A variable speed screw compressor chiller is more efficient than all variable speed centrifugal compressors with constant 85F entering condenser water temperature, see figure 3.40 below.

FIGURE 3.40 Chiller efficiency at constant 85F entering condenser water temp.

Advantages and Disadvantages of Using VFD in Screw Compressors:

Pros

• Wider modulation range than with slider system.
• Smaller compressor, lower weight, simpler design (no slider) lower compressor cost, higher reliability.
• Soft starting, low inrush current.
• No need for Power Factor correction.

Cons

• Very high inverter cost exceeds (in the lower capacity range) compressor cost savings by far.
• Increased sound levels with high speed.
• Bearing life is in direct counter-proportion to speed.
• Efficiency penalty against compressor with larger displacement by increased tip speed will result in higher throttling losses.

?Conclusion in using slide valve or VSD in screw compressor for capacity control, both modulation techniques offer specific benefits, the deciding factors for a final solution are therefore dependent on the entire system configuration, reliability, energy efficiency, investment, and maintenance cost.

Tri Rotor Screw Compressor Design

?The use of multiple male or female rotors in one screw compressor to increase capacity was proposed almost at the introduction of these machines, figure 3.41, a multirotor screw compressor with male and two female rotors layout is shown, as given by Sakun in 1960.

FIGURE 3.41 The layout of the multirotor screw compressor presented by Sakun in 1960.

The idea has not been fully commercialized until Carrier introduced the Tri-Rotor screw compressor in their 23XRV chiller line in the beginning of the 21 century.

?Tri-Rotor design allows for shorter rotors and more efficient in full load and part load than the Twin-Rotor design, the balanced rotors geometry, axial and radial thrust counter each other and the compression forces counter each other, all that will result in small bearing loads and less lubrication requirements, see figure 3.42 below.

FIGURE 3.42 A 8 to 6 lobs tri rotor set showing axial and radial forces.

?The reduction in oil into the compression process reduces the drag losses. Therefore, the Tri-Rotor design use VFD to load and unload the compressor very efficiently and increase compressor performance and reliability. Using only VFD to load and unload the compressor will enable the chiller to have a very fast response to system changes. While using a semi hermetic compressor design will allow it to be field serviced and cooled by refrigerant supply line eliminating heat rejection to mechanical room, see figure 3.43 below.

FIGURE 3.43 Tri-Rotor screw compressor with semi hermetic motor.

?The 23XRV can achieves up to 75% speed reduction and can match the load regardless of weather condition, the chiller have a wider inlet condenser water temperature range that can be from 105F down to 55F, will help the chiller to consume less KW in low load and lower entering condenser water temperatures.

?Lower Lift= Less Work= Lower KW

Oil Separation, Injection and Oil Cooling Methods in Screw Compressors

The purpose of using oil in screw compressors is:

• sealing of internal clearances between the rotors and between the rotors and the housing.
• Lubrication of bearings.
• Actuation of slide valve if used.

All the oil supplied to the compressor leaves with the refrigerant and flows to the oil separator, the separator removes the oil from the refrigerant and the refrigerant passes on to the condenser. Since the oil also absorb heat from the compression process it will be warm, therefor we need to cool it down before we send it back to the compressor, see figure 3.44 below.

FIGURE 3.44 Screw compressor oil flow and distribution.

Oil Separator: The type of oil separator used in screw compressor packages is called the coalescing type, this type of separator is much more efficient than the older style inertia separators that have been traditionally used for reciprocating compressors.

?Coalescing separators are expected to pass no liquid oil, so the only oil that escapes the separator is in vapor form, which means that the oil concentration in the refrigerant leaving the separator will be about 3 to 10 ppm.?Dirty oil will plug the coalescing elements requiring that the element be replaced. Normally the coalescing element lasts for about several years, the oil separator also serves as the oil reservoir, see figure 3.45 below.

FIGURE 3.45 Coalescing oil separator.

?The critical operating condition is at low discharge temperatures when the discharge gas will experience high mass flow rate and high specific volume, both will be contributing to high refrigerant velocities.

?Sudden drops in pressure in the separator which would generate foam should be avoided, because coalescing separators are not capable of preventing carryover of foam.

Oil Pump and Oil Flow: The injection oil flow rate should be adequate to seal the internal clearances, as well as to lubricate the moving parts and to cool the gas being compressed. On the other hand, excessive oil quantities will result in undesirable hydraulic hammer. The twin screw compressor is oil flooded compressor with around 83,000 ppm oil flow as it uses oil at the tip of the rotor for sealing. The tri rotor screw compressor is an oil reduced compressor with 166 ppm oil flow and operating in a refrigerant flooded chiller with saturated vapor feeding directly to the suction inlet, chiller generally runs with less than 15F superheat, therefore all sealing at the rotor tips is with liquid refrigerant.

There are three main methods used to achieve oil flow and injecting oil back to compressor.

• If compressor design needs full time oil pump for the compressor operation.
• Some compressor design only requires the oil pump to be operated for the start-up and shutdown. This type of oil pump is the auxiliary oil pump. The oil pump is not in operation after the compressor starts up.
• Or compressor design can start up the compressor without the auxiliary oil pump and is able to utilize the system pressure differential from the oil reservoir to other parts of the compressor for the positive oil flow, no oil pump at all.

?Oil pump is usually required if the system pressure differential is too small or when the compressor is used for booster duty or low stage application, also tri rotor compressor using a VFD for capacity control will require a full-time pump the cycle called low side oil system.

Oil Cooling Methods: In oil flooded twin screw compressor the injected oil that seals the clearances in the compressor is intimately mixed with the refrigerant undergoing compression, the refrigerant vapor becomes hot during compression and transfers some heat to the oil as it passes through the compressor therefore the oil must be cooled before reinjection.

?Various methods have been employed to cool the oil supplied to the compressor. Each of these methods of oil cooling can be categorized as either direct or indirect.

• Direct injection of liquid refrigerant into the late stages of the compressor, refrigerant will evaporate and will cool down the oil gas mixture before discharging from the compressor. The refrigerant flow control method used is a TXV the sensing bulb is on the discharge line. Advantage of this method is low cost and low maintenance. The disadvantage are capacity and power penalties, see figure 3.46 below.

FIGURE 3.46 Direct injection of liquid refrigerant into the late stages of compression.

• An indirect method by using external cooling with thermosyphon heat exchanger, the cooling medium in this method employs boiling refrigerant at saturated condensing temperature and condensing pressure. The thermosyphon principle of operation is the circulation of fluid where the motive force for fluid flow by gravitational forces and a difference in fluid densities between two vertical legs of the flow loop, and not by the addition of mechanical work like a pump, rather the transfer of heat from the oil being cooled to the refrigerant maintains the density differential necessary for flow. The basic equipment required for a thermosyphon system are, a source of refrigerant at condensing pressure and temperature located close to the compressor unit and oil cooler to minimize pressure losses in the piping, also we need an oil heat exchanger can be shell and tube were the refrigerant in the tube or can use braze plate heat exchanger. Two design are commonly used for thermosyphon oil cooling, one if the system receiver is elevated at minimum 6 feet height from the oil heat exchanger, this elevation difference is commonly referred to as (available liquid head), see figure 3.47 below, then the system high-pressure receiver can also serve as the refrigerant supply vessel for thermosyphon oil cooling heat exchanger. Liquid refrigerant draining from the condenser flows by gravity to the system receiver. The outlet connections on this vessel must be located such that the refrigerant liquid supply to the oil coolers is physically lower than the liquid supply to the system. This ensures an uninterrupted supply of refrigerant liquid to the oil cooler regardless of system demands. This can be accomplished by taking refrigerant for oil cooling off the bottom of the receiver and using a dip tube for the main system liquid line. Note that the nozzle on the receiver for the thermosyphon liquid supply, should project up into the receiver bottom floor to prevent dirt being drawn into the supply line and eventually into the oil cooler.

FIGURE 3.47 Thermosyphon oil cooling design where the level of the system receiver is above the level of the oil heat exchanger.

• Second design called Flow Through Thermosyphon Receiver, see figure 3.48 below. ?In this design all the high pressure liquid refrigerant leaving the condenser flows through the thermosyphon receiver on its way to the system receiver. Liquid refrigerant draining from the condenser flows by gravity to the thermosyphon receiver. The connections on the thermosyphon receiver (which can be either a horizontal or vertical pressure vessel) are located such that liquid refrigerant will fill this vessel up to a certain point and then overflow into the system receiver. The thermosyphon receiver must be elevated relative to the system receiver, and an equalizing line must connect these two vessels to ensure that liquid refrigerant will flow freely to the system receiver. As with the first design, the liquid level in the thermosyphon receiver must be maintained at a minimum elevation greater than 6 feet above that of the oil cooler. When the compressor is running hot oil flow through the oil heat exchanger as heat transfers from the high temperature oil to the lower temperature refrigerant the oil will cool down and the refrigerant and will start to boil. The refrigerant supply line from the thermosyphon receiver called the downcomer and it is full of refrigerant, while the riser exiting the oil cooler contains a mixture of refrigerant liquid and vapor and connected to the top of the thermosyphon receiver. Gravity causes the more dense refrigerant liquid to flow downward to the oil cooler displacing the less dense liquid vapor mixture and pushing it up the riser and back to the thermosyphon receiver. In the thermosyphon receiver, the vapor that returns from the oil cooler is separated from the liquid and vented to the inlet of the refrigerant condenser where it is once again condensed. The refrigerant in both the thermosyphon and system receivers is saturated, and its temperature is equal to the condensing temperature. Also, the refrigerant in the oil cooler is saturated at the same temperature. This refrigerant boils as it absorbs heat from the oil, but its temperature does not change. The oil therefore is cooled to within an approach temperature of the refrigerant condensing temperature. The advantages of the thermosyphon oil cooling are no capacity and power penalties, the oil heat is rejected through the refrigerant condenser and no additional heat rejection equipment needed, no risk of contamination as if the cooling medium was water or glycol, since the cooling medium is non fouling that result in higher transfer rate and longer heat exchanger life.

FIGURE 3.48 Thermosyphon oil cooling design when the level of the system receiver is at or below the level of the oil heat exchanger then requiring an additional receiver called flow through thermosyphon receiver.

• Other external cooling design is by using heat exchanger rejecting heat to a cooling medium such as water, glycol (indirect method) see figure 3.49 below. This design was the earliest and widely used concept. In warm climates cooling water could serve the heat exchanger, and this cooling water could come from a cooling tower or a closed-circuit cooler. Heat exchanger used usually is shell and tube. Maintenance is required for the water cooler due to water problems, but the advantage is that a portion of the total heat rejection from the compressor is removed by water from the oil cooler to the cooling tower or back to the source. The condenser heat rejection load is therefore smaller. In cold climates antifreeze must be circulated to avoid freezing water.

FIGURE 3.49 Oil cooling using an external heat exchanger rejecting heat to water or antifreeze.

• One other method of oil cooling is the direct injection of liquid refrigerant into the vapor refrigerant oil mixture after it leaves the compressor, see figure 3.50 below. This method extracts liquid refrigerant from the receiver and delivers it into the discharge of the compressor. Evaporation of this pumped liquid cools the refrigerant oil vapor mixture to the desired temperature before the mixture enters the separator. A temperature sensor at some point following the injection of liquid regulates a variable speed pump to adjust the flow rate of liquid. In this method and the method of direct injection of refrigerant into the late stages of the compressor the temperature of the oil separator operates at lower temperatures in contrast to the external heat exchanger and thermosyphon systems, therefore the oil separator will function slightly better that due to the viscosity as well as the vapor pressure of the oil are lowered.

FIGURE 3.50 Cooling oil by injection of liquid refrigerant into the compressor discharge line.

Using Economizer With Screw Compressor In Chiller Systems

?One of the special features of screw compressor is to allow a side port connection to the compressor casing located part way down the compression process, see figure 3.51 below.

FIGURE 3.51 Screw compressor with economizer side port.

?This feature makes economizer cycle possible for the refrigeration system to improve the system efficiency and capacity. However, screw compressor equipped with a sliding valve unloader or other device based on the same principle the economizer effectiveness quickly vanishes below full load unless the side port is part of the slide valve mechanism as in Bitzer CSH screw compressor design where the port will move with the slid and also using Vi control at part load conditions. But economizer system remains fully effective down to minimum load when capacity control is achieved exclusively by varying the compressor speed with VFD. Noted that at higher compression ratio the economizer will be more effective.

One economizer design used is by subcooling liquid from the condenser through a heat exchanger or flash tank before it goes to the evaporator, the subcooling provided by flashing liquid in the economizer cooler to intermediate pressure level, the flash gas?having absorbed heat is returned directly to the compressor at a point after suction cutoff it is mixed with gas from the suction cut-off point to produce an increase in the mass flow of refrigerant transported and compressed without either an increase in suction volume or a change in suction temperature, see figure 3.52 below.

FIGURE 3.52 Carrier 23XRV refrigerant flow schematic with flash tank economizer.

?Some manufacturers recommended to install a back pressure regulator (BPR will make sure there is always a pressure difference between the subcooled liquid in the flash vessel and the evaporator making for a steady liquid flow to the evaporator), check valve and strainer between the compressor and the economizer vessel.

?The other economizer design is using a shell and coil heat exchanger or brazed plate heat exchanger with a DX system, that design is used in low compressor capacities where the efficiency is as important as ensuring the liquid refrigerant is subcooled before going to the evaporator, see figure 3.53 below.

FIGURE 3.53 Carrier 23XRV refrigerant flow schematic with DX system economizer.