Refrigerator Troubleshooting Diagram

When a plant has been charged it must be established that the charge is in fact complete. The disappearance of bubbles in a sight glass does not necessarily mean that the evaporator is correctly flooded. Continued adding of refrigerant when bubbles show in the sight glass can also result in the system being overcharged.

Shortage of refrigerant in evaporator

A situation could arise where the sight glass shows full of liquid yet a shortage of refrigerant is evident in the evaporator. This may be due to a number of factors.

Refrigerant liquid can flash off in long liquid line runs. The ideal location for a sight glass is just before the expansion valve, although it is common practice to install them close to the condensing unit. Ideally two sight glasses should be installed, one near the condensing unit and the other before the expansion valve. This will determine whether a solid column of liquid reaches the expansion valve.

Restricted refrigerant flow to the evaporator can be caused by the following:
1. A partial blockage may occur in the filter drier, thereby creating a pressure drop in the liquid line. If the sight glass is located after the filter drier, bubbles will be evident. When a sight glass is installed before the filter drier, a pressure drop will exist just the same but bubbles will not be visible. In each case a temperature difference will exist either side of the restriction.
2. An expansion valve filter or screen may be blocked by undesirable substances circulating around the system which have passed through the filter drier, such as moisture or particles of the desiccant in the filter drier. Carbon may build up in the fine mesh of the valve screen.

3 The thermostatic expansion valve may be incorrectly adjusted or defective through partial loss of the phial charge, and therefore will not open sufficiently. A total loss of the valve phial charge will result in a complete blockage and a starved evaporator. Some expansion valves have replaceable cartridges and screens. The cartridge is stamped with an orifice size; too small an orifice can result in a starved evaporator.
4 Some plants employ evaporator defrost systems which incorporate a magnetic valve or solenoid valve. This valve, installed in the liquid line, will stop refrigerant flow to the evaporator when a defrost period is initiated by a timing device. This enables the evaporator to be evacuated of refrigerant so that, when the defrost heaters have completely cleared the evaporator of frost, an excessive build-up of pressure will be prevented during the period of continued application of heat.

Each of these conditions will result in lower than normal operating pressures.

When there is a shortage of refrigerant, pressures will be lower than normal. However, the reduction in pressure may be so slight as not to be readily detected, other than by a loss of evaporating capacity and the longer running time of the unit. If the shortage is considerable, both suction and discharge pressures will be very low; if the system temperature control consists of a thermostat only, the unit will run continuously with poor refrigerating effect. When a low pressure switch is in circuit, the compressor will short cycle (cut in and out quickly) on this control.

Likewise, any restriction in the refrigerant supply will produce the same symptoms on the low side of the system. It is possible for the compressor to operate continuously on a deep vacuum if a low pressure switch is not used. A complete blockage, preventing refrigerant from entering the evaporator, will cause a low pressure switch to stop the compressor; it will not restart because there will be insufficient pressure rise from the evaporator to actuate the
switch.

High suction and low discharge pressures

This condition is usually due to a fault within the compressor, such as broken valve reeds or incorrect seating of the valve reeds.

If the suction reeds are at fault, some of the discharged vapour will be forced back into the suction side of the compressor as the piston reaches the top of its stroke.

Faulty discharge reeds will contribute to longer running time, poor refrigeration effect and lower than normal operating head pressure. During an off cycle the discharged vapour will leak back into the cylinder(s), causing a pressure rise which will actuate a low pressure switch and quickly restart the compressor if this is the temperature control.

Figure 17 shows the passage of refrigerant through compressor valves with serviceable and defective valve reeds.

Valve reeds

Compressor valves can become distorted and fractured if liquid refrigerant is allowed to enter the cylinder; this is known as liquid slugging. For this reason service cylinders should never be inverted during the low side charging of a system.

Valve seats and reeds can become pitted if metal filings or small particles of grit are allowed to enter pipework during installation or repairs. They may circulate with the refrigerant and become trapped between the reeds and valve seats when compression takes place.

Excessive wear on pistons and cylinder walls will also contribute to a loss of compression. This is most likely to be experienced on a plant which has been in service for a long period, but could be related to lack of proper lubrication of the compressor.

A shortage of refrigerant can be the cause of oil starvation in the compressor and low pressure in the evaporator. The reduced amount of liquid refrigerant flowing through the evaporator is insufficient to move the oil and return it to the compressor, and it remains entrained in the evaporator.

This is the ratio between the suction pressure and the discharge pressure. No refrigeration compressor is 100 per cent efficient, owing to various losses. It is considered that a compressor is 80 to 85 per cent efficient with pressure ratios of 4:1 to 6:1.

With a suction pressure of 2 bar, the operating head pressure could be between 8 and 12 bar depending upon the efficiency of the compressor. If the suction pressure was 25 psig then an expected operating head of between 100 and 150 psig would be expected.

Naturally the choice of refrigerant and the system application must be taken into consideration; the values given here refer to a typical single stage reciprocating compressor.

Manufacturers of refrigeration condensing units will supply technical data for their products on demand, but this will be based on ideal operating conditions in controlled ambient temperatures. The service engineer or technician, dealing with many different types of equipment, needs a simple quick method to enable him to determine the theoretical operating head pressure of a plant so that this can be compared with the actual operating pressure to assist in fault diagnosis.

It is known that a shortage of refrigerant will produce a low suction pressure, and the fact that the compressor is doing very little work in handling vapour at a lower density will affect the pressure ratio of the compressor.

A restricted supply of refrigerant to the evaporator will not affect the operating head pressure so dramatically as an acute shortage of refrigerant, because there will be a higher pressure existing in the high side of the system, which now contains almost all of the refrigerant charge in the condenser and the receiver. It is therefore possible that only a slightly lower pressure than that to be expected will be registered in the pressure gauge.

Taking the temperature of the condensing medium

The main factor which determines the operating head pressure of a plant is the temperature of the condensing medium, which is the air passing over an air cooled condenser or the water in a water cooled system. It is important that the temperature of the condensing medium be taken accurately and from the correct location. For an air cooled unit this will be the air on to the condenser and not the air off, because the condenser will have rejected heat to the air passing over the condenser coils (see Figure 14).

A simple rule of thumb for calculating the operating head pressure of a refrigeration system is as follows:
1. Hold a thermometer in the air stream to the condenser for 2 to 3 minutes and then note the temperature.
2. Add a condensing factor of 15^o C (30^o F) to this temperature and then convert it to pressure by using a refrigerant comparator or a pressure/temperature graph, or by direct conversion from the pressure gauge on the manifold.

This will be an approximate theoretical head pressure, within 0.7 to 1 bar or 15 psig of that found under the conditions the plant is operating. An excessively high operating pressure is an indication of a condensing problem; insufficient heat is being rejected by the condenser. When taking the temperature of the water passing through a water cooled condenser, ensure that the temperature is that of the water actually entering the condenser. Some semi-hermetic motor compressors are fitted with a water coil for compressor cooling. The supply water passes through this coil before it enters the condenser coil or tubes (see Figure 15).

Water cooling of semi-hermetic motor compressor

Water cooling of open compressor

Some reciprocating open-type compressors used with water cooled systems incorporate a cylinder head cooling feature in the form of a water circuit through the compressor head. Supply water enters this before entering the condenser (see Figure 16).

A rapid indicating thermometer is recommended for taking water temperatures because the water regulating valve responds to the operating head pressure and will vary the temperature of the water as it modulates. It is also necessary, when adjusting the water regulator, to be able to take inlet and outlet temperatures quickly. Locate the thermometer probes tightly to the pipework to ensure good thermal conductivity, note the average inlet temperature and calculate as previously described.

To ensure an economical and adequate water flow through a water cooled condenser, the water regulating valve should be adjusted to provide a temperature difference of 15 to 18^o F or 7 to 9^o C between the inlet and outlet of the condenser.

The thermometer probes should be located as indicated in Figures 15 and 16.

Some examples of head pressure calculations follow. The condensing factors used are 15^o C and 30^o F

Example
An air cooled condensing unit charged with refrigerant R502 operates with the air on to the condenser at 20^o C Then 20 + 15 = 35^o C : R502 at 35^o C will register a pressure of 13.8 bar.
If the temperature of the air was lower at 15^o C then 15 + 15 = 30^o C : R502 at 30^o C will register a pressure of 12.2 bar.

Example
A water cooled condensing unit charged with refrigerant R12 operates with water entering the condenser at 50^o F Then 50 + 30 = 80^o F : RI2 at 80^o F will register a pressure of 98 psig.
If the temperature of the water is higher at 65^o F then 65 + 30 = 95^o F R12 at 95^o F will register a pressure of 115 psig.

It is important that the temperature difference (TD) between the product being stored and the refrigerant is correct for that product. Too wide a TD will result in excessive dehydration of the product. Too close a TD can result in rapid deterioration of the product; fresh meat, for example, will soon become discoloured and slimy to the touch. This section deals with storage conditions and temperature control settings, thereby giving an indication of evaporating temperatures and pressures to be expected during system operation.

Food products may be divided into four classes to provide proper storage conditions:
1. Foods which dehydrate quickly: fruits, vegetables, eggs and cheese.

2. Foods subject to sweating and some dehydration: fresh cut meats and provisions.
3. Carcase meats, chilled meats, and products not subject to excessive dehydration.
4. Products not subject to dehydration: dried fruits, tinned goods and canned/bottled beverages.

Table 2 gives the recommended TDs for these classes of product. Slight variations may be necessary to obtain ideal conditions. For better understanding, a few examples of the simple calculations needed to determine various factors are now given. These calculations are based on a system charged with refrigerant R12. See Figure 13.

Temperature difference

To determine the TD

1. Subtract the TD from the product temperature to obtain the average evaporating temperature.
2. Convert this temperature to pressure, which will be the average evaporating pressure.
3. The cut-out pressure will then be 3 psig or 0.2 bar below the average evaporating pressure.

Example
For chilled meat (product class 3), at a storage temperature of 32 to 36^o F (0 to 2.2^o C and using a forced air evaporator, the recommended TD is 12 to 15^o F (6.0 to 8.3^o C ) The average product temperature is therefore 34^o F (1.1^o C)  The cut-out pressure is 22 psig (1.5 bar), and so the average suction pressure is 22 + 3 = 25 psig (1.5 + 0.2 = 1.7 bar). This converts to an average suction temperature of 22^o F or -5^o C The TD is therefore 34 – 22 = 12^o F (1.1 – (-5) = 6.1^o C ) It can be seen that the TD is within the recommended range.

To establish the low pressure cut-out point
1. Subtract the TD from the product temperature to obtain the average evaporating temperature.
2. Convert this temperature to pressure, which will be the average evaporating pressure.
3. The cut-out pressure will then be 3 psig or 0.2 bar below the average evaporating pressure.

Example (imperial)

For a class 3 product with a storage temperature of 30^o F the recommended TD is 20 to 25^o F for a gravity coil. The average TD is 23^o F The average evaporating temperature is therefore 30- 23 = 7^o F Converted to pressure, this is 13 psig. The pressure control cut-out point will then be 13 – 3 = 10 psig. For an off-cycle defrost with a low pressure control, set the cut-in to 35 psig. This will provide an approximate coil temperature of 38^o F when the coil will be completely free of frost and ice.

Example (SI)

For a class 3 product with a storage temperature of-2^o C the recommended TD is 8^o C for a forced air evaporator. Then the average evaporating temperature is -2- 8 =-10^o C Converted to pressure, this is 1.2 bar. The cut-out pressure is established as 1.2- 0.2 = 1 bar. Temperature is -2- 8 =-10^o C  Converted to pressure, this is 1.2 bar. The cut-out pressure is established as 1.2- 0.2 = 1 bar.

When selecting cut-in and cut-out pressures to control fixture temperatures, the following information is required:
1. Dry bulb temperature of refrigerated space.
2. Relative humidity of refrigerated space.
3. Type of refrigerant used.
4. Type of evaporator.