Twin Screw Two-Phase Expanders in Large Chiller Units

I. K. Smith, N. Stosic, C. A. Aldis and A. Kovacevic

City University, London, UK

      1. SYNOPSIS
      An investigation was carried out to determine the feasibility of the use of a twin screw expander as a throttle valve replacement in a 500 ton chiller. The aim was to produce a demonstration unit with an overall machine adiabatic efficiency of not less than 70%. The efficiency target was effectively met but further analytical work is needed to predict mass flow rates reliably and hence the exact machine size required when refrigerant enters the expander as subcooled liquid.


      R134a has advantages as a refrigerant, in that it has no atmospheric ozone depleting tendencies. However, due to relatively large throttling losses, its use in chiller applications results in a reduction in the C.O.P of some 6% relative to R11. As already shown [1] this disadvantage can be largely overcome by replacing the throttle valve by a turbo expander, the output of which can be used to reduce the compressor work input. The layout of this is presented in Fig. 1.

      The use of a twin screw expander as a throttle valve replacement has been proposed or tried by several investigators [2-5] but although it has long been known that such a machine can recover power from two-phase expansion processes, the high cost of its installation hardly made its adoption worthwhile. The most significant factors responsible for this were:

      1. Poor adiabatic efficiency.
      2. High cost of construction.
      3. High cost of installation in the system.
      In all types of positive displacement machines used for expansion , the built-in volume ratio may be defined as the ratio of the displacement volume at the opening of the discharge port to the displacement volume at the point of closure of the suction port. More recent studies by the authors [5,6] have shown that low adiabatic efficiencies were mainly due to the incorrect choice of built in volume ratio which, for best results should be far less than the overall volume ratio of the actual expansion process, the latter being defined as the ratio of volume flow rate of fluid leaving the expander to volume flow rate of fluid entering it.

      Fig. 1 Chiller unit with two-phase expander as alternative to the throttle valve

      Another requirement for achieving high adiabatic efficiencies is the maintenance of near equilibrium between the liquid and vapour during the two-phase expansion process. To achieve this, the working fluid must absorb as little lubricating oil as possible and this led to the need for two features which raised the cost of construction.

      The oil lubrication system for the bearings had to be pump driven and separated from the expander working chamber by internal shaft seals of which various types have been tested. Apart from their high cost, the inclusion of the seals greatly increased the unsupported length of the rotors and led to the need for thicker rotor shafts and a larger casing which further raised the manufacturing cost. Moreover, the refrigerant in the expander working chamber has a liquid content of approximately 80% by mass and, unlike gas, this absorbs oil very readily. Hence, bearing oil leakage rates into the working chamber, which would be acceptable in gas compressors, led to unacceptably high levels of oil contamination of the working fluid in two-phase screw expanders with the seal types tried out.

      Various modes of installing the expander in the system have been proposed. These included; coupling the expander directly to the main compressor drive shaft, coupling the expander to an electrical generator and coupling the expander to a matching set of rotors in which part of the expanded vapour is recompressed [3].

                                              Fig. 2 The expander layout

      The first arrangement requires an electrical generator to be included with further loss of output through it, while both the first two proposals require a slide valve to be included in the expander in order to maintain variable flow at constant speed. The last arrangement, termed an "expressor" by the authors, could run as a self regulating device at any speed but effectively constitutes two machines in one casing.

      A final decision on which type of installation would be most suitable for installation of a screw expander in a chiller unit was not made. However, it was considered to be a worthwhile exercise to design, construct and test a simple demonstration unit without a slide valve in order to determine how well the design principles demonstrated in [7,8] would work when applied to a 500 ton chiller unit operating with R134a as the working fluid and to simplify the design as much as possible in order to minimise its cost. The target set for this exercise was to achieve adiabatic efficiencies in excess of 70%.

      The key to the expander design was the use of rotors with a profile developed by one of the authors [7,8]. These confer a number of advantages over other known types. The most important of these affecting the expander are large cross sectional flow area with small blow hole area, small torque transmission through the female rotor and involute shape on the contact band which is located close to the pitch circle diameter. These factors led to a compact machine with rotors between which there was estimated to be almost pure rolling and low contact stresses.

      It was decided that with these rotors it would be possible to dispense with the timing gear and rely on the presence of liquid refrigerant to act as a coolant should any heat be generated by rotor contact. This was first checked by using the same profile rotors in an oil injected compressor in which water was used as the injected fluid [9]. This ran successfully for 150 hours without any detectable rotor wear at higher pressure differences than were required for the expander.

      Fig 3. A view of the expander, as installed in the test rig

      Advances in rolling element bearing design [10] led to the decision to use liquid refrigerant as the bearing coolant and thereby completely eliminate oil from the expander system. Internal shaft seals then would not be required.

      These assumptions led to the design of a machine with rotors of 5/6 configuration, a male rotor diameter of 127.5mm and an L/D ratio of 1.65. The built in volume ratio was set at 2.85:1 in order to achieve an overall volume ratio of 12.3:1. Only one external shaft seal is required, of the type used in refrigeration compressors, and as shown in Fig. 2, the layout is very simple.

      The expander was built and installed in a rig specially designed for two-phase expansion testing, the arrangement of which is shown in Fig. 4 and a more detailed description given in [5]. Essentially, it removes the need for the large power inputs associated with a vapour compression cycle system compressor by using a high boiling point working fluid (R113). The refrigerant is pressurised by a liquid feed pump, heated to its boiling point only, without evaporation, expanded as a two-phase fluid and then condensed at above atmospheric temperature. A parallel loop enables the fluid to be circulated with the expander bypassed and this is used both for setting up the required temperature, pressure and flow conditions and as a safety feature in the event of expander failure. A view of the expander, as installed in the test rig is given in Fig 3.

      The rig was originally designed and constructed in 1984 and before use on the expander described in this paper, it was extensively refurbished, especially with regard to the data acquisition, display and processing. This included the use of four piezzo resistive pressure transducers located within the expander to record and display how the pressure changes within the expander as a function of rotational angle. The mode of screen data presentation employed is shown in Fig 5.

      Fig. 4 City University expander test rig

      A key requirement of the experimental programme was to simulate an R134a expansion under chiller design conditions as closely as possible. This involved the working fluid being admitted to the machine as subcooled liquid; a process not previously attempted. The most rigorous replication when using R113, would be to keep the expander inlet and exit pressures the same as in a chiller operating on R134a with an equal amount of subcooling at the inlet. At the design point, this corresponded to R113 entering the expander at a pressure of 8.45 bar and a temperature of 126oC and leaving at 3.76 bar with a corresponding saturation temperature of 93oC. With R113, the volume expansion ratio would then be 20% greater and the mass flow 15% higher than with R134a, the larger mass flow being mainly due to the higher liquid density of R113. Up to the time of writing, the analytical model for predicting the expander performance would not solve if subcooling at the inlet was assumed. This is because it is based on the non-steady flow energy equation applied to a variable volume and small changes in volume induce instantaneous phase changes. Computer simulations of the expander performance with both fluids were therefore carried out on the assumption of saturated liquid at the expander inlet. The results showed that under these conditions, the power output with R113 would be approximately 12% greater and the adiabatic efficiency about 2% less than when using R134a. These differences were considered to be sufficiently small to make the R113 test results meaningful, bearing in mind that the installation of a two-phase expander of 70% adiabatic efficiency would increase the overall COP of the chiller unit by approximately 7%.

      In view of the very small temperature and specific enthalpy changes associated with the planned tests, every effort was made to ensure the accuracy of the pressure, temperature, mass flow, speed and torque measurements and all instruments were repeatedly calibrated against instruments with recognised standards certificates. Despite this, it was almost impossible to obtain temperature measurements at the exit without some small degree of superheat compared to saturated values at the measured pressure. It was therefore concluded that the fluid was leaving the expander in a non-equilibrium condition with a small amount of liquid superheat.

      Fig. 5 A screen print

      Nonetheless, in estimating the adiabatic efficiency of the expander it was assumed that the ideal condition of the fluid at exit was with the liquid and vapour in thermodynamic equilibrium since this is the most rigorous criterion applicable.

      At the time of writing this paper, the test results were still incomplete and there appear to be some anomalies in the results at higher mass flow rates. Nonetheless, some indication of the expander performance is given in Figs 6 to 9 in which all points shown correspond to test conditions with subcooled liquid admission. In Figs 7 and 9 the abscissa was taken as specific enthalpy drop rather than pressure ratio. This was used to give a better appreciation of the very small enthalpy changes associated with two-phase expansion.


      Fig. 6 Power output variation with rotational speed


      Fig. 7 Power output variation with specific enthalpy drop

      An error analysis was carried out on the test results. These showed the following:

      Measurement errors of expander inlet pressure had a negligible effect since the fluid was subcooled. Inlet temperature measurement errors produce an adiabatic efficiency error of 5% per 1oC. Exit pressure measurement errors produce an adiabatic efficiency error of 5% per 0.1Bar. Errors in adiabatic efficiency estimates due to
      errors in fluid flow and torque measurements would each be of the order of 1%.

      Fig. 8 Adiabatic efficiency variation with rotational speed

      Overall it is estimated that the uncertainty in the derived results is of the order of ±3% or approximately ±2 percentage points at an estimated adiabatic efficiency of 70%. This is too small to affect the results shown in Figs. 6 to 9 radically.

      Graphical display of results permits an output variable to be shown as a function of two input variables. This is sufficient for unambiguous presentation of single-phase expansion processes of fixed pressure ratio. In the case of two-phase expansion there is an additional variable of inlet fluid quality, which is not displayed. Also it was extremely difficult to control all the variables while maintaining a constant pressure ratio. These factors are the main cause of the apparent scatter of the test points.

      As may be seen, a substantial number of test points were obtained at adiabatic efficiencies of over 70%. These amounted to approximately 28% of the total. To the best of the authors’ knowledge these are the highest values ever achieved in any type of two-phase expander and in this respect the performance predictions were realised. The corresponding mass flow measurements were 20-40% less than those predicted from the assumption of saturated liquid admission. Test results of two-phase expansion through a nozzle [1] showed that in steady flow, subcooling increased the mass flow rate and this is in agreement with qualitative reasoning on how the machine swallowing capacity should be affected. A convincing explanation for this apparent anomaly in screw expander performance has not yet been found and investigations are continuing on how to resolve it as well as finding a means of analysing the transition from subcooled liquid to two-phase mixture during discrete system changes.

      Fig. 9 Adiabatic efficiency variation with specific enthalpy drop

      After completion of the construction of the expander, cost estimates were obtained for the manufacture of such a machine in small batches.

      The experimental programme described in this report has shown that though reliable modelling of the two-phase expansion of subcooled liquids has not yet been achieved, it is possible to produce screw expanders of simple design and low manufacturing cost with adiabatic efficiencies in excess of 70% at power outputs of the order of 10-20 kW. These are suitable for use as throttle valve replacements in large chiller units. Preliminary cost studies indicate that energy savings resulting from their installation would yield simple repayment times of less than one year.


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      2. Smith, I. K. and Aldis, C. A. Lysholm screw expanders in place of throttle valves in large heat pump and refrigeration systems, Proc Meeting of Commission E2, Stockholm, August 29-31,1990.
      3. Smith, I. K. and Stosic, N. R. The Expressor: An efficiency boost to vapour compression systems by power recovery from the throttling process. AES-vol.34, Heat pump and refrigeration systems design, Analysis and applications, ASME 1995, pp. 173-181.
      4. Taniguchi, H, Giedt, W. H, Kudo, K, Kasahara, K, Ohta, J and Kawamura, K. Energy-conserving heat pump-boiler systems for district heating. Proceedings of Eighteenth Intersociety Energy Conversion Conference, 1983, v4, pp. 1862-1868.
      5. Smith, I. K. Stosic, N and Aldis, C. A. Development of the trilateral flash cycle system Part 3: the design of high efficiency two-phase screw expanders. Proc Instn Mech Engrs, Part A, 1996, 210(A2), 75-93.
      6. Smith, I. K and Stosic, N. R. US Pat 5,833,446, November 10th 1998.
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      8. Stosic, N, Smith, I. K, Kovacevic, A and Aldis, C. A. The design of a twin-screw compressor based on a new rotor profile. Jrnl of Engng Design, v.8, n.4, 1997, pp389-399
      9. Stosic, N, Smith, I. K, Brasz, J. J and Sishtla, V. The performance of a screw compressor with involute contact rotors in a low viscosity gas-liquid mixture environment. VDI Berichte NR. 1391, 1998, pp 279-292.
      10. Jacobson, B. Ball-bearing lubrication in refrigeration compressors. Proc 1996 International Compressor Conference at Purdue, v1, Purdue University, West Lafayette, IND, USA, July 23-26, 1996, pp 103-108.